Spark ignition engine control device

ABSTRACT

For the purpose of improving the fuel efficiency by lean combustion and enhancing the fuel efficiency improvement effects by performing compression ignition efficiently in some cylinders, a multi-cylinder spark ignition engine is constructed such that exhaust gas, that is exhausted from preceding cylinders  2 A,  2 D on the exhaust stroke side among pairs of cylinders whose exhaust stroke and intake stroke overlap in a low load, low rotational speed region, is directly introduced through an inter-cylinder gas passage  22  into following cylinders  2 B,  2 C on the intake stroke side and only gas exhausted from the following cylinders  2 B,  2 C is fed to an exhaust passage  20,  which is provided with a three-way catalyst  24.  Combustion controller is provided that controls the combustion of each of the cylinders such that combustion is conducted by forced ignition in a condition in which the air/fuel ratio is a lean air/fuel ratio which is larger by a prescribed amount than the stoichiometric air/fuel ratio in the preceding cylinders  2 A,  2 D and, in the following cylinders  2 B,  2 C, fuel is supplied to burnt gas of lean air/fuel ratio introduced from the preceding cylinders  2 A,  2 D and combustion is conducted by compression ignition.

TECHNICAL FIELD

The present invention relates to a control device for a spark ignitionengine and in more detail relates to a device that controls thecombustion condition in the cylinders of a multi-cylinder engine inorder to improve fuel consumption and reduce emissions.

BACKGROUND ART

Techniques are previously known for improving fuel consumption in sparkignition engines by performing combustion under lean air/fuel ratioconditions, in which the air/fuel ratio of the mixture in the cylindersis larger than the stoichiometric air/fuel ratio (or theoreticalair/fuel ration). For example, as illustrated in Laid-open JapanesePatent Application No. H. 10-274085, a technique is known in which aninjection valve that injects fuel directly into the combustion chamberis provided and super-lean combustion is produced by conductingstratified charge combustion in the low rotational speed low load regionetc. Specifically, such stratified charge combustion consists inaltering the composition ratio of the mixture in the vicinity of thespark plug in the ignition period by injecting fuel in the compressionstroke, while controlling the rate of air intake and rate of fuelinjection such as to produce a condition in the combustion chamber as awhole that is much leaner than the stoichiometric air/fuel ratio, andperforming combustion with forced ignition by the spark plug in thiscondition.

When super-lean combustion is performed by stratified charge combustionas described above, thermal efficiency is improved and the air intakerate becomes large, reducing the intake negative pressure and therebygreatly improving fuel consumption. Also, in such a super-leanstratified charge combustion condition, even if some of the air that ispresent in excess is replaced by EGR, fully satisfactory combustion isstill achieved, so a comparatively large amount of EGR may be employedand this is thereby beneficial in lowering NOx etc. Thus, even thoughthis large amount of EGR is introduced, the benefit of a lowered pumpingloss is still obtained and thermal efficiency is also increased comparedwith ordinary combustion in which the air intake rate and EGR rate arecontrolled without layering; the benefit of improved fuel consumption isthereby obtained.

However, when stratified charge combustion is performed, although, asthe air/fuel ratio is made leaner, improved fuel consumption is obtainedup to a certain point, if the mixture becomes leaner than a certaindegree, the combustion rate becomes too low, with the result that thecombustion occurring in the vicinity of the final period does notcontribute to work, so, contrariwise, fuel consumption tends todeteriorate. Thus, there were limits to the extent to which fuelconsumption improvement could be achieved by increasing leanness instratified charge combustion.

Compression ignition has been studied as another technique for improvingfuel consumption. This compression ignition consists in self-ignition offuel at high temperature and high pressure in a combustion chamber inthe latter period of the compression stroke, in the same way as in thecase of a diesel engine. If such compression ignition is performed, evenunder conditions of a super-lean air/fuel ratio or conditions ofintroduction of a large amount of EGR, combustion occurs at oncethroughout the entire combustion chamber. Slow combustion, which doesnot contribute to work, is thereby avoided, which is beneficial inimproving fuel consumption.

However, in an ordinary spark ignition engine (gasoline engine), forcedignition for combustion is necessary and the temperature and pressurewithin the combustion chamber in the vicinity of the top dead center incompression are not elevated to a sufficient degree to producecompression ignition; thus special expedients must be adopted if thetemperature or pressure in the combustion chamber is to be raised to theconsiderable degree necessary to achieve compression ignition. However,in a conventional spark ignition engine, it is difficult to raise thetemperature or pressure in the combustion chamber to such an extent asto produce compression ignition in the low load region where fuelconsumption improvement is required while yet preventing knocking in thehigh load region, so implementation of such a technique has not beenachieved.

In view of the aforementioned problems, the present invention provides acontrol device for a spark ignition engine wherein the benefit ofimproved fuel consumption is produced by lean combustion and, inaddition, the benefit of improved fuel consumption is increased byeffectively performing compression ignition in a portion of thecylinders.

DISCLOSURE OF THE INVENTION

According to the present invention, in a multi-cylinder spark ignitionengine wherein the cylinders are arranged to perform a cycle comprisingintake, compression, expansion and exhaustion strokes with prescribedphase differences, at least in a low load, low rotational speed region,a gas flow path is constituted in a two-cylinder connected conditionsuch that burnt gas exhausted from a preceding cylinder (or a leadingcylinder) which is a cylinder on the exhaust stroke side in a pair ofcylinders whose exhaustion stroke and intake stroke overlap is arrangedto be directly introduced into a following cylinder which is a cylinderon the intake stroke side through an inter-cylinder gas passage and gasexhausted from this following cylinder is arranged to be fed to anexhaust passage; and combustion controller is provided that controlscombustion in each cylinder such that at least in part of the operatingregion of the operating region in which said two-cylinder connectedcondition is produced, combustion is performed by forced ignition insaid preceding cylinder in a condition with a lean air/fuel ratio whichis larger by a prescribed amount than the stoichiometric air/fuel ratiowhile fuel is supplied in an amount corresponding to the followingcylinder to the burnt gas generated by combustion in this precedingcylinder, and combustion is performed by compression self-ignition inthe following cylinder.

If such a construction is adopted, at least in the low load, lowrotational speed region thermal efficiency is raised by lean combustionin the preceding cylinder and pumping loss is diminished, making itpossible to obtain a considerable fuel costs improvement effect. Also,in the case of the following cylinder, combustion is conducted bysupplying fuel to the burnt gas of lean air/fuel ratio introduced fromthe preceding cylinder, so, since this burnt gas is at high temperature,the temperature in the combustion chamber rises to such a degree thatcompression ignition can be achieved in the final period of thecompression stroke and compression ignition is therefore performed.Thus, by introducing burnt gas, the same condition is produced in thefollowing cylinder as if a large amount of EGR were introduced but,since combustion is performed rapidly by compression ignition even insuch a condition, the combustion contributes efficiently to the work andfuel costs are greatly improved by this and by the decreased pumpinglosses.

In a device according to the present invention, preferably the air/fuelratio of the following cylinder when in said two-cylinder connectedcondition is made to be at or below the stoichiometric air/fuel ratioand a three-way catalyst or oxidation catalyst is provided in theexhaust passage connected with this following cylinder.

In this way, although combustion is conducted in the preceding cylinderwith a lean air/fuel ratio, since gas of below the stoichiometricair/fuel ratio is introduced into the exhaust passage, a lean NO_(x)catalyst is unnecessary and problems such as compromise of the fuelcosts improvement effect or sulfur poisoning, due to temporaryenrichment of the air/fuel ratio, are obviated.

Also, if a fuel injection valve is provided that injects fuel directlyinto the cylinder in respect of said preceding cylinder and, when insaid two-cylinder connected condition, fuel is injected in thecompression stroke from said fuel injection valve and stratified chargecombustion is performed by forced ignition while keeping a lean air/fuelratio in the preceding cylinder, combustion with a super-lean air/fuelratio becomes possible by stratified charge combustion in the precedingcylinder, increasing the fuel costs improvement effect.

If the air/fuel ratio of the preceding cylinder when in saidtwo-cylinder connected condition is twice or more the stoichiometricair/fuel ratio, an ample fuel costs improvement effect can be obtainedby lean combustion in the preceding cylinder and burnt gas containing alarge amount of excess air is delivered to the following cylinder fromthe preceding cylinder, which is beneficial for combustion in thefollowing cylinder.

Also, preferably the air/fuel ratio of the following cylinder when insaid two-cylinder connected condition is a lean air/fuel ratio greaterthan the stoichiometric air/fuel ratio.

In this way, combustion is performed rapidly by compression ignition inthe following cylinder while keeping a lean air/fuel ratio, so theamount of NO_(x) generated is diminished and the fuel costs improvementeffect is increased.

Preferably, when in said two-cylinder connected condition, uniformcombustion is performed by injecting fuel in the following cylinder inthe intake stroke. If this is done, burnt gas of lean air/fuel ratio andfuel are uniformly mixed in the following cylinder, so that combustioncan be satisfactorily performed by compression self-ignition.

Also, preferably, there is provided flow path changeover means wherebyin a high load, high rotational speed operating region, the flow pathsof new air and gas are changed over such that the intake port andexhaust port of each of the cylinders are made to be independent, sothat new air is introduced into the intake port of each cylinder from anintake passage and exhaust gas exhausted from the exhaust port of eachcylinder is fed to said exhaust passage and combustion controller isarranged to make the air/fuel ratio of each of the cylinders thestoichiometric air/fuel ratio or less than this and to cause combustionto be performed by forced ignition in each of the cylinders in said highload, high rotational speed operating region.

In this way, it is possible to prevent the heat load on the followingcylinder becoming excessively high and to ensure output performance inthe high load, high rotational speed operating region.

If this is done, preferably, in said preceding cylinder, there areprovided an intake port that communicates with said intake passage, afirst exhaust port that communicates with said exhaust passage and asecond exhaust port that communicates with the inter-cylinder gaspassage and, in said following cylinder there are provided a firstintake port that communicates with said intake passage, a second intakeport that communicates with said inter-cylinder gas passage and anexhaust port that communicates with said exhaust passage and, as saidflow path changeover means, there are provided a valve deactivatingmechanism that changes over the operating condition and deactivatedcondition respectively of the first and second exhaust valves that openand close the first and second exhaust ports of said preceding cylinderand of the first and second intake valves that open and close the firstand second intake ports of the following cylinder; and valve stopmechanism controller that, in a low load, low rotational speed region,puts said first exhaust valve and said first intake valve in deactivatedcondition and puts said second exhaust valve and said second intakevalve in operating condition and, in a high load, high rotational speedoperating condition, puts said first exhaust valve and said first intakevalve in operating condition and said second exhaust valve and saidsecond intake valve in deactivated condition.

In this way, by control of the valve deactivation mechanism, changeoverof flow path can easily be effected in accordance with the operatingcondition in the low load, low rotational speed region or on the highload side/high rotational speed side.

Also, in a device according to the present invention, it is effective toprovide combustion condition controller that exercise control such thatthe control mode whereby combustion is performed in said two-cylinderconnected condition is the special operating mode and, in at least partof the operating region of the region that is put in the specialoperating mode, the fuel supply rates in respect of both the leading andfollowing cylinders are controlled such that the fuel supply rate in thepreceding cylinder is greater, while the air/fuel ratio duringcombustion in said following cylinder is substantially thestoichiometric air/fuel ratio, thereby making the air/fuel ratio whencombustion is conducted in the preceding cylinder a value of less thantwice the stoichiometric air/fuel ratio and conducting combustion in thepreceding cylinder by forced ignition and conducting combustion in thefollowing cylinder by compression self-ignition.

If this is done, since the gas that is exhausted to the exhaust passagefrom the following cylinder is of the stoichiometric air/fuel ratio,cleansing of the exhaust gas can be achieved fully satisfactorily simplyby a three-way catalyst and, by making the air/fuel ratio of thepreceding cylinder a value of less than twice the stoichiometricair/fuel ratio by making the fuel supply rate in respect of thepreceding cylinder larger, the temperature of the gas that is introducedinto the following cylinder from the preceding cylinder is increased,thereby improving the self-ignition capability of the following cylinderand increasing the amount of burnt gas constituents corresponding to EGRin this gas, etc, and so improving the knocking suppression effect.

Preferably, in said special operating mode, in the intermediate speedregion of the operating region in which the following cylinder is madeto perform compression self-ignition, the air/fuel ratio when conductingcombustion in the preceding cylinder is made to be a value ofsubstantially twice the stoichiometric air/fuel ratio, or more thanthis.

If this is done, the fuel costs improvement effect is increased in theintermediate speed region of the operating region in which the followingcylinder is made to perform compression self-ignition in the specialoperating mode.

If, in this way, in said special operating mode, in the operating regionon the low speed side of the intermediate speed region of the operatingregion in which the following cylinder is made to perform compressionself-ignition, the air/fuel ratio when conducting combustion in thepreceding cylinder is made to be a value of less than twice thestoichiometric air/fuel ratio, the self-ignition capability in theoperating region on low speed side of the intermediate region of theoperating region in which the following cylinder is made to performcompression self-ignition in the special operating mode is improved.

Furthermore, if, in said special operating mode, in the operating regionon the high speed side of the intermediate speed region of the operatingregion in which the following cylinder is made to perform compressionself-ignition, the air/fuel ratio when conducting combustion in thepreceding cylinder is made to be a value of less than twice thestoichiometric air/fuel ratio, occurrence of knocking is suppressed inthe operating region on the high speed side of the intermediate speedregion of the operating region in which the following cylinder is madeto perform compression self-ignition in the special operating mode.

Also, preferably, in said special operating mode, in the intermediateload region of the operating region in which the following cylinder ismade to perform compression self-ignition, the air/fuel ratio whenconducting combustion in the preceding cylinder is made to be a value ofsubstantially twice the stoichiometric air/fuel ratio, or more thanthis.

In this way, the fuel costs improvement effect in the intermediateregion of the operating region in which the following cylinder is madeto perform compression self-ignition in the special operating mode isincreased.

Also, preferably, in said special operating mode, in the intermediatespeed/intermediate load region of the operating region in which thefollowing cylinder is made to perform compression self-ignition, theair/fuel ratio when conducting combustion in the preceding cylinder ismade to be a value of substantially twice the stoichiometric air/fuelratio, or more than this.

In this way, the fuel costs improvement effect in the intermediatespeed/intermediate load region of the operating region in which thefollowing cylinder is made to perform compression self-ignition in thespecial operating mode is increased.

Also, in said special operating mode, in the operating region in whichthe following cylinder is made to perform compression self-ignition, theair/fuel ratio when conducting combustion in the preceding cylinder maysuitably be made smaller as the load becomes lower.

In this way, the tendency for compression self-ignition to become moredifficult as the load becomes lower in the operating region in which thefollowing cylinder is made to perform compression self-ignition in thespecial operating mode is compensated.

Suitably also, when the engine temperature is low, in the entireoperating region in which the following cylinder is made to performcompression self-ignition in said special operating mode, the air/fuelratio when conducting combustion in the preceding cylinder is made to beless than twice the stoichiometric air/fuel ratio.

In this way, compression self-ignition can be achieved even at lowengine temperature.

It is also effective to provide a fuel injection controller that causes,in an operating region in which the following cylinder is made toperform compression self-ignition, in the case where there is anoperating condition in which knocking is likely to occur, the combustioncontrol means to retard the injection time of the fuel with respect tothe following cylinder compared with the case where there is anoperating condition in which knocking is unlikely to occur.

If this is done, in the operating region in which the following cylinderis made to perform compression self-ignition, in an operating conditionin which knocking is likely to occur, the fuel injection time in regardto the following cylinder is relatively retarded, thereby suppressingactivation of the mixture and so effectively preventing occurrence ofknocking caused by the self-ignition capability of the mixture becomingtoo high. Also, in an operating region in which the following cylinderis made to perform compression self-ignition, in an operating conditionin which knocking is unlikely to occur, the injection time of the fuelin respect of the following cylinder is relatively advanced, sooccurrence of misfiring in the following cylinder due to activation ofthe mixture is effectively prevented and an improvement in thermalefficiency can be achieved by maintaining the compression self-ignitioncapability.

Suitably also, in an operating region in which the following cylinder ismade to perform compression self-ignition in said special operatingmode, in an operating condition in which knocking is likely to occur,the injection time of the fuel in respect of the following cylinder isset more on the retarded side of the compression stroke as thelikelihood of knocking increases.

If this is done, occurrence of knocking in an operating region whereknocking is liable to occur due to high temperature of the combustionchamber of the following cylinder in the region where the followingcylinder is made to perform compression self-ignition can be effectivelyprevented by suppressing activation of the mixture to an appropriateextent by relatively retarding the injection time of the fuel in regardto the following cylinder so that it is set in the latter half of thecompression stroke.

Suitably also, in an operating region in which the following cylinder ismade to perform compression self-ignition in said special operatingmode, in an operating condition in which knocking is likely to occur,injection of fuel into the following cylinder is performed in dividedfashion and the latter injection time of the fuel in this dividedinjection is set in the latter half of the compression.

If this is done, in an operating region in which the following cylinderis made to perform compression self-ignition, activation of the mixtureis suppressed to an appropriate extent, making it possible toeffectively prevent occurrence of knocking while preventing occurrenceof misfiring.

Suitably also, in a region in which the following cylinder is made toperform compression self-ignition, the likelihood of occurrence ofknocking or the intensity of knocking is ascertained and the latterinjection time in said divided fuel injection is retarded so as toapproach more closely the compression top dead center as the likelihoodof occurrence of this knocking or the intensity of knocking increases.

If this is done, in an operating region in which the following cylinderis made to perform compression self-ignition, activation of the mixtureis suppressed to an appropriate extent, making it possible toeffectively prevent occurrence of knocking while preventing occurrenceof misfiring.

Suitably also, in a region in which the following cylinder is made toperform compression self-ignition, in an operating condition in whichknocking is likely to occur, injection of fuel into the followingcylinder is performed in divided fashion and the latter injection rateof the fuel in this divided injection is set to a larger value than theformer injection rate.

If this is done, in an operating region in which the following cylinderis made to perform compression self-ignition, occurrence of knocking iseffectively prevented while preventing occurrence of misfiring.

Preferably, if this is done, in a region in which the following cylinderis made to perform compression self-ignition, the likelihood ofoccurrence of knocking is ascertained and the ratio of the latterinjection period rate with respect to the total injection rate of fuelinjected in the following cylinder is changed so as to be increased asthe likelihood of occurrence of such knocking becomes higher.

If this is done, in an operating region in which the following cylinderis made to perform compression self-ignition, occurrence of knocking ismore effectively prevented by further suppressing activation of themixture.

Suitably also, in a region in which the following cylinder is made toperform compression self-ignition, when the engine is in an operatingregion on the high load side, a condition in which knocking is likely tooccur is identified.

If this is done, in an operating region in which the following cylinderis made to perform compression self-ignition, the injection time of thefuel in regard to the following cylinder is appropriately controlled.

Suitably also, if fuel of low octane value is employed, in a region inwhich the following cylinder is made to perform compressionself-ignition, this is identified as a condition in which knocking islikely to occur.

In this way, in a region in which the following cylinder is made toperform compression self-ignition, the injection time of the fuel withregard to the following cylinder is appropriately controlled.

If this is done, suitably there is provided swirl generating means thatgenerates swirl such that a strong intensity of turbulence is maintainedin the latter half of the compression stroke in a region in which thefollowing cylinder is made to perform compression self-ignition, in anoperating condition in which knocking is likely to occur.

If this is done, in an operating region in which the following cylinderis made to perform compression self-ignition, the effect of improvingthe combustibility by maintaining a strong intensity of turbulence inthe latter half of the compression stroke and the effect of suppressingknocking by retarding the injection time of the fuel in regard to thefollowing cylinder so that it approaches more closely to the compressiontop dead center can be combined.

Suitably also, swirl is generated in the combustion chamber by directingthe tip portion of the inter-cylinder gas passage in the cylindertangential direction of the following cylinder in plan view andintroducing burnt gas into the following cylinder from saidinter-cylinder gas passage in the intake stroke of the followingcylinder.

In this way, a fully satisfactory combustible condition of the followingcylinder is maintained by forming swirl such that a strong intensity ofturbulence is maintained in the latter half of the compression stroke,by introducing burnt gas exhausted from the preceding cylinder into thefollowing cylinder with the inter-cylinder gas passage in a conductingcondition in the following cylinder intake stroke in a region in whichthe following cylinder is made to perform compression self-ignition.

Also, it is effective if a combustion control means of the deviceaccording to the present invention effects control, in at least part ofthe operating region that has been put into said special operating mode,combustion is conducted by compression self-ignition in the followingcylinder, and, in a high load region in the region in which thiscompression self-ignition is performed, the air/fuel ratio of thepreceding cylinder is made relatively rich compared with the region onthe low load side of this and a new air introduction intake valve thatintroduces new air into the following cylinder is opened, therebyintroducing new air into the following cylinder in addition to the burntgas that is fed from said preceding cylinder.

If this is done, if, in a region on the high load side in the regionwhere the following cylinder is made to perform compressionself-ignition, the air/fuel ratio of the preceding cylinder is madecomparatively rich and the oxygen concentration in the burnt gasintroduced into the following cylinder correspondingly falls, new air isthen introduced into the following cylinder by opening the new airintroduction intake valve, so that compression self-ignition in thefollowing cylinder is thereby appropriately conducted by eliminating thedeficiency of new air in the following cylinder and occurrence ofknocking is effectively prevented by increasing the amount of burnt gasconstituents introduced into the following cylinder and engine output isthereby guaranteed.

Preferably, in a region on the low load side in the operating region inwhich the following cylinder is made to perform compressionself-ignition in said special operating mode, the new air introductionintake valve is maintained in closed condition; and, in a region on thehigh load side in said compression self-ignition region, the new airintroduction intake valve is opened in the vicinity of the intake topdead center of the following cylinder and is closed during the course ofthe intake stroke of the following cylinder.

If this is done, in the low load region in the region in which thefollowing cylinder is made to perform compression self-ignition, theoxygen concentration in the burnt gas that is introduced into thisfollowing cylinder being maintained at a high value by making theair/fuel ratio of the preceding cylinder comparatively lean, theair/fuel ratio in the following cylinder can be prevented from becominglean by keeping the new air introduction intake valve in a closedcondition. Also, by opening the new air introduction intake valve in thevicinity of the intake of top dead center of the following cylinder in aregion on the high load side in the compression self-ignition region,efficient introduction of new air into the following cylinder and bystopping introduction of new air by closing this during the intakestroke of the following cylinder, smooth introduction of the burnt gasthat is fed from the preceding cylinder into the following cylinder areachieved.

Preferably, in a region on the high load side in the operating region inwhich the following cylinder is made to perform compressionself-ignition in said special operating mode, the burnt gas introductionvalve of the following cylinder is opened during the course of theintake stroke and the new air introduction intake valve is opened priorto the opening time of this burnt gas introduction valve.

If this is done, in a region on the high load side in the operatingregion in which the following cylinder is made to perform compressionself-ignition, new air is efficiently introduced into the followingcylinder and the burnt gas introduction valve is then closed during thecourse of the intake stroke of the following cylinder, thereby ensuringthat the burnt gas that is fed from the preceding cylinder is introducedinto the following cylinder.

Preferably, in a region on the high load side in the operating region inwhich the following cylinder is made to perform compressionself-ignition, control is exercised such as to increase the ratio of thenew air intake rate with respect to the total gas rate introduced intothe following cylinder, in response to enrichment of the air/fuel ratioof the preceding cylinder, compared with a region on the low load sidethereof.

Also, in a region on the high load side in the region in which thefollowing cylinder is made to perform compression self-ignition, theratio of the rate of introduction of new air with respect to the totalgas rate introduced into the following cylinder is controlled so as tobe raised in response to the enrichment of the air/fuel ratio of thepreceding cylinder compared with a region on the low load side thereof.

In this way, in a region on the high load side in the operating regionin which the following cylinder is made to perform compressionself-ignition, if the air/fuel ratio of the preceding cylinder iscomparatively enriched and the oxygen concentration in the gasintroduced into the following cylinder corresponding falls, the ratio ofthe rate of new air introduction with respect to the total gas rateintroduced into the following cylinder is raised, so that deficiency ofnew air in the following cylinder is efficiently eliminated andcompression self-ignition in the following cylinder therebyappropriately performed and occurrence of knocking effectively preventedby suppressing rise in temperature in the following cylinder.

Also, preferably, at least in a region in which the following cylinderis made to perform compression self-ignition, the air/fuel ratio of thefollowing cylinder is controlled such that the oxygen concentration inthe exhaust gas that is exhausted from the following cylinder is a valuecorresponding to the combustion condition of the stoichiometric air/fuelratio.

In this way, at least in the region in which the following cylinder ismade to perform compression self-ignition, albeit combustion isconducted with a lean air/fuel ratio in the preceding cylinder, onlyburnt gas of the following cylinder that has been burnt with thestoichiometric air/fuel ratio is fed to the exhaust passage.

Also, in a device according to the present invention, it is effective ifthe combustion control means effects control such that the totalinjection rate of fuel injected into the two cylinders i.e. saidpreceding cylinder and following cylinder is increased in response toincrease in engine load; and control is exercised such that in saidfollowing cylinder, combustion is conducted by compression self-ignitionin at least part of the operating region in which said special operatingmode is produced and, in said preceding cylinder, control is exercisedsuch that stratified charge lean combustion is conducted with theinjected fuel put in a stratified charge condition in anintermediate/low load region of the operating region in whichcompression self-ignition of said following cylinder is performed; andcontrol is exercised such that, on the high load side of the operatingregion in which this stratified charge lean combustion is conducted,uniform lean combustion is conducted in a condition with the injectedfuel uniformly dispersed.

In this way, control is exercised so as to change over the combustioncondition in the preceding cylinder in accordance with the load regionof the engine, so improvement in fuel costs can be appropriatelyachieved while yet effectively preventing knocking. For example, in anoperating region in which combustion self-ignition is performed in thefollowing cylinder, in an intermediate/low load region in which thetotal injection rate of fuel is relatively small, further improvement infuel costs can be achieved while maintaining combustion stability withstratified charge lean combustion. In contrast, on the high load side ofthis intermediate/low load region, by making the air/fuel ratio arelatively smaller value as the total fuel injection rate is increasedand conducting uniform lean combustion, a lower combustion temperaturecan be achieved than in the case of a uniform lean condition with thesame air/fuel ratio condition and occurrence of knocking in thefollowing cylinder can be prevented by suppressing the rise oftemperature of the burnt gas introduced into the following cylinder, orlowering this, and the region in which compression self-ignition of thefollowing cylinder is feasible can be expanded. As a result, the fuelcosts improvement effect can be further increased.

Preferably, in the operating region on the high load side wherecombustion is conducted in a uniform lean condition in the precedingcylinder, the air/fuel ratio of said preceding cylinder is made to be avalue of substantially twice the stoichiometric air/fuel ratio, or avalue smaller than this.

If this is done, it might be feared that misfiring could occur if theair/fuel ratio under uniform lean conditions in the preceding cylinderbecomes higher than prescribed, but combustion is stabilized by makingthe air/fuel ratio substantially twice the stoichiometric air/fuelratio, or a value smaller than this, and the rise in temperature of theburnt gas introduced into the following cylinder is suppressed.

Also, preferably, in a low load operating region of the intermediate/lowload operating region in which stratified charge lean combustion isconducted in said preceding cylinder, the air/fuel ratio of saidpreceding cylinder is made to be a value of substantially twice thestoichiometric air/fuel ratio, or a value smaller than this.

If this is done, in an intermediate/low load region where the fuelinjection rate is small, the temperature of the burnt gas introducedinto the following cylinder is raised by the stratified charge leancombustion, so the range in which compression self-ignition in thefollowing cylinder is feasible can be expanded on the low load side.

Suitably also, in a low load operating region of the intermediate/lowload operating region in which stratified charge lean combustion isconducted in said preceding cylinder, if compression self-ignition insaid following cylinder is difficult, control is exercised such that theair/fuel ratio of said preceding cylinder is made to be substantiallytwice the stoichiometric air/fuel ratio or a value smaller than this andthe combustion mode in the preceding cylinder is shifted from thestratified charge lean condition to said uniform lean condition and theignition mode in said following cylinder is shifted from compressionself-ignition to forced ignition.

If this is done, in cases where, if the engine has not been sufficientlywarmed up etc, compression self-ignition in the following cylinder isdifficult, the air/fuel ratio of the preceding cylinder is reduced,thereby raising the temperature of the burnt gas introduced into thefollowing cylinder; also, by using uniform lean combustion an adverseeffect on fuel costs concomitant with this air/fuel ratio condition issuppressed and the shift from forced ignition of the following cylinderto compression self-ignition can be made earlier.

Also, in a device according to the present invention, it is effective ifthe flow paths of intake and exhaust are arranged to be capable of beingchanged over, these flow paths being arranged be capable of beingchanged over in operating mode between an ordinary operating mode inwhich each of the cylinders are put in an independent condition in whichcombustion is conducted respectively independently in each of thecylinders and a special operating mode in which combustion is conductedin said two-cylinder connected condition; comprising: first fuelinjection means that supplies fuel independently in respect of each ofthe cylinders when in said ordinary operating mode; second fuelinjection means whereby it is made possible to supply fuel in an amountcorresponding to that of the following cylinder in respect of said burntgas prior to introduction thereof into the following cylinder aftercompletion of combustion in said preceding cylinder, when in saidspecial operating mode; and the combustion controller that, when in saidordinary operating mode, conduct combustion in a condition with theair/fuel ratio in each cylinder made to be the stoichiometric air/fuelratio by supplying fuel by said first fuel injection means and, when inthe special operating mode, conduct combustion in the preceding cylinderby forced ignition in a condition with the air/fuel ratio made to be alean air/fuel ratio greater by a prescribed amount than thestoichiometric air/fuel ratio, by supplying fuel by said first fuelinjection means and that control combustion such as to conductcombustion by compression self-ignition in each cylinder by introducinggas in a condition of the stoichiometric air/fuel ratio by supplyingfuel in the following cylinder to said burnt gas by said second fuelinjection means.

In this way, in the preceding cylinder, thermal efficiency is raised bylean combustion and pumping loss is diminished and, in the followingcylinder, compression self-ignition is performed by supplying fuel tothe burnt gas from the preceding cylinder; in this way, combustion isconducted rapidly so that the combustion contributes efficiently to thework i.e. the benefits of efficient combustion and diminished pumpingloss are obtained and, as a result, fuel costs are considerablyimproved. In addition, after completion of combustion in the precedingcylinder, fuel in an amount corresponding to that for the followingcylinder is supplied thereto in regard to the following cylinder and isthereby introduced into the following cylinder in a condition in whichthe mixture is thoroughly mixed with the high temperature burnt gas; asa result, the capability for self-ignition in the following cylinder isimproved.

In this case, preferably, said first fuel injection means is arrangedsuch as to inject fuel directly into the combustion chamber in respectof said preceding cylinder and the first fuel injection means of saidpreceding cylinder also serves as said second fuel injection means, whenin said special operating mode, by constituting said fuel controllersuch that supply of fuel for the following cylinder in respect of saidburnt gas is performed by said first fuel injection means of thepreceding cylinder during the exhaustion stroke of this cylinder.

In this way, since the fuel is supplied in respect of the burnt gas atan early stage, the burnt gas and the mixture are introduced into thefollowing cylinder in a condition in which they are more effectivelymixed, so the self-ignition capability in the following cylinder iseffectively increased. Also, by supplying fuel in an amountcorresponding to that for the following cylinder by the first fuelinjection means of the preceding cylinder, this basic construction canbe utilized without modification for example in a direct injection typeengine provided with an injector (fuel injection means) for in-cylinderinjection into each cylinder, making it possible to apply the presentinvention in a way that is consistent with general objectives.

If this is done, said first fuel injection means may be arranged suchthat fuel is injected into an intake passage in respect of saidfollowing cylinder. That is, since, in regard to the preceding cylinder,the first fuel injection means may be arranged so as to be capable ofinjection into the cylinder, in regard to the following cylinder, thefirst fuel injection means may be arranged so as to inject fuel into anintake passage.

Suitably also, said second fuel injection means is provided at somepoint along said inter-cylinder gas passage and fuel is supplied therebyin an amount corresponding to that of the following cylinder in respectof said burnt gas after exhaustion from the preceding cylinder prior tointroduction thereof into the following cylinder.

In this way, better activation is achieved by supplying fuel in anamount corresponding to that of the following cylinder in the specialoperating mode to the burnt gas flowing through the inter-cylinder gaspassage.

Suitably also, said fuel controller, when in said special operatingmode, is constituted such as to be capable of changing over the fuelinjection mode between the first injection mode in which combustion isconducted by compression ignition by supplying fuel in an amountcorresponding to the following cylinder in respect of said burnt gas bythe first fuel injection means of this following cylinder afterintroduction of burnt gas into the following cylinder from saidpreceding cylinder; and a second injection mode in which combustion isconducted by compression self-ignition by supplying fuel in an amountcorresponding to the following cylinder by said second fuel injectionmeans in respect of said burnt gas prior to introduction thereof intothe following cylinder after completion of combustion in said precedingcylinder and is constituted such as to determine the degree ofcapability of self-ignition of the following cylinder from informationrelating to the operating condition and to be capable of changing overthe fuel injection mode in accordance with the results of thisdetermination.

In this way, the self-ignition capability of the following cylinder inthe special operating mode is increased by changing over the fuelinjection mode in accordance with operating condition.

If this is done, said combustion controller may be constituted such asto put said injection mode into the second injection mode when in anoperating condition wherein the degree of capability for self-ignitionof the following cylinder is low.

In this way, combustion stability in the following cylinder is improvedby raising the self-ignition capability in the following cylinder due toa mixing effect of the burnt gas and the mixture, when in an operatingcondition in which the degree of self-ignition capability of thefollowing cylinder is low.

Suitably also, said fuel injection means is constituted such as todetermine that the operating condition is one in which the degree ofcapability for self-ignition is low if the cylinder temperature is belowa specified temperature after warming up operation.

If this is done, the capability for self-ignition in the followingcylinder is raised by putting the fuel injection mode in the secondinjection mode on ascertaining that there is an operating condition inwhich the self-ignition capability is low when the cylinder temperatureis below the specified temperature and the cylinder temperature is lowafter warm-up operation.

Suitably also, said combustion controller is constituted such as todetermine that the operating condition is one in which the degree ofcapability for self-ignition is low when in a very low load region.

If this is done, the capability for self-ignition in the followingcylinder is raised by putting the fuel injection mode in the secondinjection mode on ascertaining that there is an operating condition inwhich the self-ignition capability is low when, in a very low loadregion, the fuel injection rate is low.

It is also effective if, in a device according to the present invention,there are provided a preceding cylinder intake valve whereby new air isintroduced into said preceding cylinder and a burnt gas introductionvalve whereby burnt gas is introduced into said following cylinder fromsaid inter-cylinder gas passage when in said two-cylinder connectedcondition and in at least a prescribed region on the low load side ofsaid operating region that is in a two-cylinder connected condition, theinterval between the intake stroke bottom dead center of said followingcylinder and the closure time of said burnt gas introduction valve isset to be shorter than the interval between the intake stroke bottomdead center of said preceding cylinder and the closure time of saidpreceding cylinder intake valve.

If this is done, at least in a prescribed region on the low load side inthe operating region which is in a two-cylinder connected condition, theclosure time of the burnt gas introduction valve of the followingcylinder is closed earlier than in the case of new air intake in thepreceding cylinder, so the effective compression ratio of the followingcylinder is increased, facilitating self-ignition are due to the rise incylinder temperature. Consequently, even in a low load region where thecapability for self-ignition is low, due to difficulty in raising thecylinder temperature, the capability for self-ignition is improved andcombustion by compression self-ignition can be further expanded into thelow load region, so promoting further improvement in fuel costs andexhaust cleansing.

Preferably, in this case, there is provided a following cylinder exhaustvalve that exhausts exhaust gas of said following cylinder and in atleast a prescribed region on the low load side of said operating regionthat is in a two-cylinder connected condition, the opening time of saidburnt gas introduction valve is set to be the intake stroke top deadcenter of said following cylinder, while said following cylinder exhaustvalve is open until the top dead center of the exhaust stroke of saidfollowing cylinder.

If this is done, the overlap of valve opening of the following cylinderexhaust valve and the burnt gas introduction valve is shortened, withthe result that so-called “blow through”, in which burnt gas that isintroduced into the following cylinder is directly exhausted to theexhaust passage through the exhaust valve of the following cylinder isprevented and the effective compression ratio of the following cylinderis increased, increasing the self-ignition capability and promotingfurther improvement in fuel costs and exhaust cleansing.

Preferably also, in a prescribed region on the high load side of saidoperating region that is in a two-cylinder connected condition, theclosure time of said burnt gas introduction valve is set on the delayedside from this time when in the prescribed region on the low load side.

If this is done, when, conversely, there is a risk of occurrence ofabnormal combustion such as knocking with unnecessarily high cylindertemperature the cylinder temperature may be lowered by decreasing theeffective compression ratio of the following cylinder by delaying theclosure time of the burnt gas introduction valve; abnormal combustioncan thereby be prevented and the operating region in which combustioncan be conducted by compression self-ignition thereby expanded.

Preferably, also, in a prescribed region on the high load, highrotational speed side of said operating region that is in a two-cylinderconnected condition, the closure time of said burnt gas introductionvalve is set on the delayed side from this time when in the prescribedregion on the low load, low rotational speed side.

If this is done, control may be performed taking into account speed ofrotation also; more precise and appropriate control of the compressionself-ignition capability can thereby be achieved.

Preferably also, a burnt gas exhaust valve is provided that exhaustsburnt gas of said preceding cylinder to said inter-cylinder gas passagewhen in said two-cylinder connected condition and in the operatingregion that is in said two-cylinder connected condition, the closuretime of said burnt gas exhaust valve is set on the advancing side of theclosure time of said burnt gas introduction valve and while maintainingthe open period of said burnt gas exhaust valve and the open period ofsaid burnt gas introduction valve at fixed prescribed values, theopening time of said burnt gas exhaust valve and the opening time ofsaid burnt gas introduction valve are set so as to vary forwards andbackwards in accordance with engine load while maintaining thedifference of these times fixed.

In this way, when under comparatively low load, the effectivecompression ratio of the following cylinder is increased by advancingthe closure time of the burnt gas introduction valve and since theclosure time of the burnt gas exhaust valve is then on the advancingside of the closure time of the burnt gas introduction valve, the amountof burnt gas left behind in the preceding cylinder is increased,increasing the cylinder temperature, with the result that the cylindertemperature of the following cylinder rises due to rise in temperatureof the burnt gas; the compression self-ignition region can thereby beexpanded on the low load side.

In contrast, under comparatively high load, the opening period as awhole is retarded and the closure of the burnt gas introduction valve isretarded, thereby diminishing the effective compression ratio of thefollowing cylinder and lowering the burnt gas temperature by decreasinginternal EGR of the preceding cylinder; abnormal combustion of thefollowing cylinder is thus prevented, thereby making it possible toexpand the compression self-ignition region on the high load side. As aresult, improvement of fuel costs and exhaust gas cleansing can befurther promoted.

It should be noted that, since the time difference of these valveopening times is arranged to be maintained constant, with the openingperiod of the burnt gas exhaust valve and the opening period of theburnt gas introduction valve maintained at fixed prescribed values, inan engine of a construction in which the opening/closure times of eachof the cylinders are uniquely set by the shape of the cams relating toopening/closure of the respective valves, there is no need for camchangeover etc and the same cams can always be employed as the camsrelating to the respective valves. Variation of the valve opening timescan be achieved by varying the phase of the crankshaft and the camshaftthat rotates integrally with the cams that relate to opening/closure ofthese valves, so the construction of the engine in question can besimplified compared with the case where the respective cams arecontrolled independently and this therefore makes possible reductions insize, weight and costs.

Also, in a device according to the present invention, it is effectiveif, when in said two-cylinder connected condition there are provided apreceding cylinder intake valve that introduces new air into saidpreceding cylinder and a burnt gas introduction valve that introducesburnt gas into said following cylinder from said inter-cylinder gaspassage and in at least a prescribed region on the low load side of theoperating region that is in said two-cylinder connected condition, theopen period of said burnt gas introduction valve is set so as to beshorter than the open period of said preceding cylinder intake valve.

In this way, when in a running condition under comparatively low load,the open period of the burnt gas introduction valve of the followingcylinder can be made shorter than the open period of the precedingcylinder intake valve, so the closure time of the burnt gas introductionvalve can be made relatively earlier, thereby making it possible toincrease the effective compression ratio of the following cylinder.Consequently, even in a low load region where the capability forcompression ignition is low due to difficulty in raising the cylindertemperature, the self-ignition capability can be improved by increasingthe effective compression ratio of the following cylinder, therebyfurther improving fuel costs and promoting exhaust gas cleansing.

Also, in a device according to the present invention, it is effective ifthe device is constituted such that, in said two-cylinder connectedcondition and in a prescribed region on the comparatively low load sideof the operating region in which combustion is conducted by compressionself-ignition in the following cylinder, combustion is conducted bycompression self-ignition in said preceding cylinder while increasingthe amount of internal EGR of said preceding cylinder and the internalEGR ratio is decreased with increase in load.

In this way, in a prescribed region of comparatively low load in theoperating region in which combustion is conducted by compressionself-ignition and the following cylinder, in a condition in which hightemperature burnt gas is left behind in the preceding cylinder, this iscarried over to the next intake stroke and compression stroke, so thecylinder temperature rises, facilitating compression self-ignition sothat combustion is conducted by compression self-ignition; in this way,high thermal efficiency and an NOx suppression effect can be obtained inthe same way as in the case of the following cylinder, conferring thebenefit of further improvement in fuel costs and an exhaust gascleansing effect.

It should be noted that, in addition to diminution of the rate of newair by increase in the amount of internal EGR in the preceding cylinder,the injection rate becomes comparatively low due to combustion with alean air/fuel ratio; however, the region in which the internal EGRincreases in the preceding cylinder is made to be a prescribed region onthe comparatively low load side, so the required output can be obtainedeven with a comparatively small fuel injection rate.

Preferably, if this is done, in part or all of the operating region inwhich combustion is conducted by compression self-ignition in both saidpreceding cylinder and said following cylinder, the closure time of theburnt gas exhaust valve that exhausts burnt gas to said inter-cylindergas passage in the exhaust stroke provided in said preceding cylinder isset earlier than the top dead center of the exhaust stroke of saidpreceding cylinder.

In this way, in a condition in which a large amount of burnt gas is leftbehind in the cylinder, this can be carried over into the next intakestroke and compression stroke.

If this is done, preferably, in part or all of the operating region inwhich combustion is conducted by compression self-ignition in both saidpreceding cylinder and said following cylinder, said combustioncontroller sets the injection time of fuel into said preceding cylinderlater than the closure time of said burnt gas exhaust valve and in thevicinity of the top dead center of the exhaust stroke.

If this is done, since fuel is injected into the preceding cylinder inwhich a large amount of burnt gas is left behind, activation of the fuelby the high temperature of this burnt gas can be achieved and, inaddition, activation is promoted since fuel injection is performedearly, in the vicinity of the top dead center of the exhaust stroke; thecompression self-ignition capability can thereby be improved. It shouldbe noted that there is no possibility of the injected fuel beingdirectly exhausted, since the fuel injection is performed after closureof the burnt gas exhaust valve.

Preferably also, in part or all of the operating region in whichcombustion is conducted by compression self-ignition in both saidpreceding cylinder and said following cylinder, said combustioncontroller exercises control such that the air/fuel ratio in saidfollowing cylinder is substantially a lean air/fuel ratio.

If this done, combustion is conducted with a lean air/fuel ratio notonly in the preceding cylinder but also in the following cylinder, sothermal efficiency can be further raised and generation of NOxsuppressed; in addition, generation of NOx is further suppressed bycompression self-ignition, enabling the exhaust cleansing performance tobe further improved.

Preferably, if this is done, the catalyst for exhaust gas cleansingprovided in said exhaust passage consists solely of a three-way catalystor solely of a three-way catalyst and oxidation catalyst.

In this way, generation of NOx is suppressed to a fully satisfactoryextent due to the effect of the lean air/fuel ratio in both thepreceding cylinder and following cylinder and due to the compressionself-ignition, so fully satisfactory exhaust gas cleansing performancecan be obtained with only a three-way catalyst or only a three-waycatalyst and an oxidation catalyst; a lean NOx catalyst is thereforeunnecessary.

It is also effective if there is provided a burnt gas introduction valvethat introduces burnt gas from said inter-cylinder gas passage in theintake stroke when in said two-cylinder connected condition, provided insaid following cylinder, and a following cylinder intake valve thatintroduces new air in the intake stroke when in said two-cylinderconnected condition, provided in said following cylinder; and in partall of the operating region in which combustion is conducted bycompression self-ignition in both said preceding cylinder and saidfollowing cylinder, the opening time of said burnt gas introductionvalve is set on the delayed side of the top dead center of the intakestroke of this following cylinder and said following cylinder intakevalve is arranged to open earlier than the opening time of said burntgas introduction valve.

In this way, since, apart from burnt gas, new air is also introducedinto the following cylinder from the following cylinder intake valve,even if, due to increase in the amount of internal EGR in the precedingcylinder, there is little oxygen in the burnt gas that is introducedinto the following cylinder, the output generated in the followingcylinder can be increased. Also, since the limit of increase of theamount of internal EGR in the preceding cylinder is increased, theregion in which compression self-ignition in the preceding cylinder isfeasible is expanded.

It should be noted that, since the burnt gas introduction valve isopened later than the following cylinder intake valve, direct exhaustionof the burnt gas through the following cylinder intake valve isprevented.

Preferably, if this is done, said preceding cylinder is of the longstroke type and comprises a preceding cylinder intake valve thatintroduces new air in the intake stroke when in said two-cylinderconnected condition and in part or all of the operating region in whichcombustion is conducted by compression self-ignition in both saidpreceding cylinder and said following cylinder the closure time of saidburnt gas exhaust valve and said burnt gas introduction valve is set onthe delayed side of the top dead center of the exhaust stroke of saidpreceding cylinder and the opening time of said preceding cylinderintake valve is set earlier than the top dead center of the intakestroke of this preceding cylinder.

If this is done, the mixed flow rate of the new air and the burnt gas isincreased by increasing the opening overlap period of the burnt gasexhaust valve and the preceding cylinder intake valve in the precedingcylinder, thereby enabling the amount of internal EGR to be increased.

Also, even if the opening overlap period is increased, interferencethereof can be prevented by shortening the period for which the pistonis in the vicinity of the top dead center, by adopting a long stroketype cylinder.

Also, it is effective if there is provided a supercharger thatsupercharges the intake in respect of said preceding cylinder and inpart or all of the operating region in which combustion is conducted bycompression self-ignition in at least said preceding cylinder and saidfollowing cylinder, supercharging is performed using said supercharger.

In this way, the rate of introduction of new air in the precedingcylinder is increased and, concomitantly, the amount of internal EGR canalso be increased and the intake temperature further raised bysupercharging, thereby making it possible to expand the operating regionin which combustion is conducted by compression self-ignition in thepreceding cylinder and so making possible a further improvement in fuelcosts.

Preferably also, in a prescribed region on the comparatively high loadside of said operating region in which combustion is conducted bycompression self-ignition in said following cylinder, said combustioncontroller conducts combustion by forced ignition in said precedingcylinder and, when in an operating region in which combustion isconducted by compression self-ignition in both said preceding cylinderand said following cylinder, performs setting such that the air/fuelratio of said preceding cylinder is substantially larger.

In this way, in a comparatively high load region, the rate ofintroduction of new air can be increased by decreasing the internal EGRratio in the preceding cylinder, so that, even if the cylindertemperature drops, combustion is conducted by forced ignition, so stablecombustion can be achieved.

Also, if combustion in the preceding cylinder is conducted bycompression self-ignition, even if the air/fuel ratio is a lean air/fuelratio, it is necessary that this should be kept comparatively on therich side but there is no particular need for this after changeover toforced ignition, so by setting a large air/fuel ratio in the precedingcylinder and setting the following cylinder comparatively on the richside to perform combustion using compression self-ignition, which has acorrespondingly better thermal efficiency, further improvement in fuelcosts can be achieved.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a diagrammatic plan view of an entire engine comprising adevice according to an embodiment of the present invention;

FIG. 2 is a diagrammatic cross-sectional view of a main engine unit etc;

FIG. 3 is a block diagram of a control system;

FIG. 4 is a diagram showing an example of operating region setting forcontrol in accordance with the operating condition;

FIG. 5 is a view showing the exhaust stroke, intake stroke, fuelinjection period and ignition period etc of each cylinder;

FIG. 6 is a diagram showing a flow path for substantially new air andgas during low load, low rotational speed operation;

FIG. 7 is a diagram showing a flow path for substantially new air andgas in an operating region on the high load, high/low^(i) rotationalspeed side;

FIG. 8 is a diagrammatic plan view of an entire engine showing amodified example in which the catalyst etc provided in the exhaustpassage is changed from that shown in FIG. 1;

FIG. 9 is a diagram showing an example of operating region setting forexercising control in accordance with the operating condition inaccordance with another embodiment;

FIG. 10 is a diagram showing a second example in regard to operatingregion setting for exercising control in accordance with the operatingcondition;

FIG. 11 is a diagram showing a third example in regard to operatingregion setting for exercising control in accordance with the operatingcondition;

FIG. 12 is a diagram showing a fourth example in regard to operatingregion setting for exercising control in accordance with the operatingcondition;

FIG. 13 is a diagram showing a fifth example in regard to operatingregion setting for exercising control in accordance with the operatingcondition;

FIG. 14 is a diagram showing an example of setting of operating regionfor exercising control in accordance with the operating condition inaccordance with yet a further embodiment;

FIG. 15 is a diagram showing the combustion cycle of a precedingcylinder and a following cylinder;

FIG. 16 is a diagram showing a further example of the combustion cycleof a preceding cylinder and a following cylinder;

FIG. 17 is a diagram showing the specific construction of swirlgenerating means;

FIG. 18 is a diagram showing an example of the setting of the operatingregion for exercising control in accordance with the operating conditionin accordance with yet a further embodiment;

FIG. 19 is a diagram showing the combustion cycle and valve openingtiming of a preceding cylinder and a following cylinder;

FIG. 20 is a block diagram of a control system showing yet a furtherembodiment;

FIG. 21 is a diagram showing an example of the setting of the operatingregion for exercising control in accordance with the operating conditionby means of a device according to the embodiment shown in FIG. 20;

FIG. 22 is a view showing the relationship between burnt gas temperatureand air/fuel ratio under the same load in stratified charge leancombustion and uniform lean combustion;

FIG. 23 is a view showing the exhaust stroke, intake stroke, fuelinjection period and ignition period etc of each cylinder in the case ofa special operating mode in which a preceding cylinder is made toperform uniform lean combustion while a following cylinder is made toperform forced ignition;

FIG. 24 is a view showing the exhaust stroke, intake stroke, fuelinjection period and ignition period etc of each cylinder in the case ofa special operating mode in which a preceding cylinder is made toperform uniform lean combustion while a following cylinder is made toperform compression self-ignition;

FIG. 25 is a view showing the relationship between load and air/fuelratio in a preceding cylinder;

FIG. 26 is a block diagram of a control system showing yet a furtherembodiment;

FIG. 27 is a view showing the exhaust stroke, intake stroke, fuelinjection period and ignition period etc of each cylinder;

FIG. 28 is a diagrammatic plan view showing a modified example of anengine;

FIG. 29 is a view showing the exhaust stroke, intake stroke, fuelinjection period and ignition period etc of each cylinder in the case ofthe modified example shown in FIG. 28;

FIG. 30 is a diagrammatic plan view of an entire engine according to yeta further embodiment;

FIG. 31 is a block diagram of a control system of the same embodiment;

FIG. 32 is a diagram showing an example of the setting of the operatingregion for exercising control in accordance with the operatingcondition;

FIG. 33 is a diagram showing the opening/closing times of anintake/exhaust valve in a special operating mode, (a) showing the caseof comparatively low load, low rotational speed and (b) showing in likemanner the case of comparatively high load, high rotational speed;

FIG. 34 is a diagram showing the opening/closing times of anintake/exhaust valve in the ordinary operating mode;

FIG. 35 is a partial perspective view showing a cam changeover mechanismemployed in a yet a further embodiment;

FIG. 36 is a plunger action diagram given in explanation of three typesof cam changeover mechanism;

FIG. 37 is a plunger action diagram given in explanation of two types ofcam changeover mechanism;

FIG. 38 is a block diagram of a control system in an embodimentemploying a cam changeover mechanism;

FIG. 39 is a diagram showing the opening/closing times of anintake/exhaust valve in a special operating mode, (a) showing the caseof comparatively low load, low rotational speed and (b) showing in likemanner the case of comparatively high load, high rotational speed;

FIG. 40 is a diagrammatic plan view of an entire engine according to yeta further embodiment;

FIG. 41 is a diagrammatic cross-sectional view of the main engine unitetc according to this embodiment;

FIG. 42 is a partial perspective view showing a cam changeover mechanismemployed in this embodiment;

FIG. 43 is a plunger action diagram for a cam changeover mechanism;

FIG. 44 is a block diagram of a control system;

FIG. 45 is a diagram showing an example of the setting of the operatingregion for exercising control in accordance with the operatingcondition;

FIG. 46 is a diagram showing the opening/closing times of anintake/exhaust valve in a special operating mode, (a) showing the caseof comparatively low load and (b) showing in like manner the case ofintermediate load;

FIG. 47 is a diagram showing the opening/closing times of anintake/exhaust valve in a special operating mode, showing the case ofcomparatively high load;

FIG. 48 is a diagram showing the opening/closing times of anintake/exhaust valve in the ordinary operating mode;

FIG. 49 is a diagram showing the opening/closing times of anintake/exhaust valve in a special operating mode according to a secondexample of the control of intake/exhaust etc using a device as shown inFIG. 40 to FIG. 44, (a) showing the case of comparatively low load and(b) showing in like manner the case of a comparatively high load;

FIG. 50 is a diagram showing the opening/closing times of anintake/exhaust valve in a special operating mode according to a thirdexample of the control of intake/exhaust etc using a device as shown inFIG. 40 to FIG. 44, (a) showing the case of comparatively low load and(b) showing in like manner the case of a comparatively high load;

FIG. 51 is a diagram showing the opening/closing times of anintake/exhaust valve in a special operating mode according to a fourthexample of the control of intake/exhaust etc using a device as shown inFIG. 40 to FIG. 44, (a) showing the case of comparatively low load and(b) showing in like manner the case of an intermediate load; and

FIG. 52 is a diagrammatic plan view of an entire engine showing yet afurther embodiment.

BEST MODE FOR CARRYING OUT THE INVENTION

Embodiments of the invention are described below with reference to thedrawings.

FIG. 1 shows diagrammatically the construction of an engine according toan embodiment of the present invention and FIG. 2 shows diagrammaticallythe construction of one cylinder of a main engine body 1 and anintake/exhaust valve etc provided in respect of this. In these Figures,the main engine body 1 comprises a plurality of cylinders; in theembodiment shown, there are four cylinders 2A to 2D. A piston 3 isinserted into each of the cylinders 2A to 2D, a combustion chamber 4being formed above the piston 3.

A spark plug 7 is provided at the top of the combustion chamber 4 ofeach cylinder 2, the tip of this plug facing the interior of thecombustion chamber 4. An ignition circuit 8 capable of controlling theignition time by electronic control is connected with this spark plug 7.

A fuel ignition valve 9 that directly injects fuel into the combustionchamber 4 is provided at the side of the combustion chamber 4. Thisfuel-injection valve 9 incorporates a needle valve and a solenoid, notshown, and is constructed such that, by input of a pulse signal, to bedescribed, drive is effected to open the valve for a time correspondingto the pulse width at the time of this pulse input so as to inject fuelin an amount depending on the valve opening period. The fuel supplysystem is constituted such that fuel is supplied through a fuel supplypassage etc by a fuel pump, outside the Figure, to this fuel-injectionvalve 9 and that a fuel pressure that is higher than the pressure withinthe combustion chamber is applied in the compression stroke.

Also, by opening of intake ports 11, 11 a, 11 b and exhaust ports 12, 12a, 12 b with respect to the combustion chamber 4 of the respectivecylinders 2A to 2D, an air intake passage 15 and exhaust passage 20 etcare connected with these ports and these ports are arranged to be openedand closed by means of intake valves 31, 31 a, 31 b and exhaust valves32, 32 a, 32 b.

The cylinders are arranged to perform a cycle comprising intake,compression, expansion and exhaust strokes with prescribed phasedifferences. In the case of a four-cylinder engine, if the cylinders areidentified as a first cylinder 2A, a second cylinder 2B, a thirdcylinder 2C and four-cylinder 2D from one end in the direction of thearrangement of the cylinders, as shown in FIG. 5, the aforesaid cycle isarranged to be performed with crank angle phase differences of 180° ineach case, in the order: first cylinder 2A, third cylinder 2C, fourthcylinder 2D and second cylinder 2B. In FIG. 5, EX is the exhaust strokeand IN is the intake stroke; also, F represents fuel injection and Srepresents forced ignition; the asterisks in the Figure indicate thatcompression ignition is performed.

In a pair of cylinders whose exhaust stroke and intake stroke overlap,there is provided an inter-cylinder gas passage 22 such that burnt gascan be directly fed from the cylinder which is on the exhaust strokeside when the exhaust stroke and intake stroke overlap (in thespecification, this will be termed the preceding cylinder) to thecylinder on the side of the intake stroke (in the specification thiswill be termed the following cylinder). In the four-cylinder engine ofthis embodiment, as shown in FIG. 5, the exhaust stroke (EX) of thefirst cylinder 2A and the intake stroke (IN) of the second cylinder 2Boverlap and the exhaust stroke (EX) of the fourth cylinder 2D and theintake stroke (IN) of the third cylinder 2C overlap, so the firstcylinder 2A and second cylinder 2B and the fourth cylinder 2D and thirdcylinder 2C respectively constitute pairs, the first cylinder 2A andfourth cylinder 2D being preceding cylinders, while the second cylinder2B and third cylinder 2C are following cylinders.

Specifically, the intake/exhaust ports of each cylinder and the intakepassage, exhaust passage and inter-cylinder passage connected theretoare constructed as follows.

The preceding cylinders i.e. the first cylinder 2A and fourth cylinder2D are respectively provided with an intake port 11 for introduction ofnew air, a first exhaust port 12 a for delivering burnt gas (exhaustgas) to the exhaust passage, and a second exhaust port 12 b for feedingout burnt gas to the following cylinders. Also, the following cylindersi.e. the second cylinder 2B and third cylinder 2C are respectivelyprovided with a first intake port 11 a for introducing new air, a secondintake port 11 b for introducing burnt gas from the preceding cylindersand an exhaust port 32 for delivering burnt gas to the exhaust passage.

In the example shown in FIG. 1, two per cylinder of the intake port 11in the first and fourth cylinders 2A and 2D and the first intake port 11a in the second and third cylinders 2B and 2C are provided in parallelon the left-hand side of the combustion chamber, and a first exhaustport. 12 a and second exhaust port 12 b in the first and fourthcylinders 2A and 2D and a second intake port 11 b and exhaust port 12 bin the second and third cylinders 2B and 2C are provided in parallel onthe right-hand side of the combustion chamber.

The downstream end of the branch intake passage 16 for each cylinder inthe intake passage 15 is connected with the intake port 11 in the firstand fourth cylinders 2A and 2D and with the first intake port 11 a inthe second and third cylinders 2B and 2C. In the vicinity of thedownstream end of each branch intake passage 16 there are providedmultiple linked throttle valves 17 that are mutually linked by means ofa common shaft, these multiple linked throttle valve 17 being driven bymeans of an actuator 18 in response to a control signal, to adjust theair intake rate. An airflow sensor 19 is provided that detects the airintake flow rate in the common air intake passage upstream of themerging section in the intake passage 15.

The upstream end of a branched exhaust passage 21 for each cylinder inthe exhaust passage 20 is connected to the first exhaust ports 12 a inthe first and fourth cylinders 2A and 2D and to the exhaust ports 12 inthe second and third cylinders 2B and 2C. Also, respectiveinter-cylinder gas passages 22 are provided between the first cylinder2A and second cylinder 2B and between the third cylinder 2C and fourthcylinder 2D. The upstream end of the inter-cylinder gas passage 22 isconnected with the second exhaust ports 12 b of the first and fourthcylinders 2A and 2D, which are the preceding cylinders, and thedownstream end of the inter-cylinder gas passage 22 is connected withthe second intake port 11 b of the second and third cylinders 2B and 2C,which are the following cylinders.

The aforementioned inter-cylinder gas passages 22 are comparativelyshort passages that make connections between mutually adjacentcylinders, so that the amount of heat radiated whilst the gas exhaustedfrom the preceding cylinders passes through these passages 22 is kept toa comparatively low level.

An O₂ sensor 23 is provided that detects the air/fuel ratio by detectingthe oxygen concentration in the exhaust gas in the merging sectiondownstream of the branched exhaust passage 21 in the exhaust passage 20.In addition, an exhaust gas cleaning catalyst is provided in the exhaustpassage 21 downstream of the O₂ sensor 23; in this embodiment, a leanNOx catalyst 24A and three-way catalyst 24B are provided. The lean NOxcatalyst 24A has an NOx cleansing ability even at lean air/fuel ratiosand comprises for example an occlusion-type lean NOx catalyst thatadsorbs NOx in an excess oxygen atmosphere and performs release andreduction of NOx in an atmosphere of lowered oxygen concentration. Also,as is generally known, the three-way catalyst 24B is a catalyst thatshows a high cleansing ability in respect of HC, CO and NOx when theair/fuel ratio of the exhaust gas is in the vicinity of thestoichiometric air/fuel ratio (i.e. the excess air ratio λ is λ=1).

The intake/exhaust valves that open/close the intake/exhaust ports ofthe cylinders and the valve actuation mechanisms of these are asfollows.

The intake port 11, first exhaust port 12 a and second exhaust port 12 bat the first and fourth cylinders 2A and 2D are respectively providedwith an intake valve 31, first exhaust valve 32 a and second exhaustvalve 32 b and the first intake port 11 a, second intake port 11 b andexhaust port 12 at the second and third cylinders 2A and 2B arerespectively provided with a first intake valve 31 a, second intakevalve 31 b and exhaust valve 32. Also, in order that the intake strokeand exhaust stroke of the cylinders should be performed with theprescribed phase differences described above, these intake/exhaustvalves are driven so as to open/close with prescribed timings by meansof a valve actuation mechanism comprising respective camshafts 33 and 34etc.

In addition, of these intake/exhaust valves, the first exhaust valve 32a, second exhaust valve 32 b, first intake valve 31 a and second intakevalve 31 b are provided with a valve deactivating mechanism 35 thatchanges over the valves between an actuated condition and a deactivatedcondition. This valve deactivating mechanism 35 is not shown in detail,since it is previously known, but is for example such that a hydraulicchamber capable of supplying/draining hydraulic fluid in respect of atappet interposed between a valve shaft and cam of camshafts 33 and 34is provided wherein, in a condition in which hydraulic fluid is suppliedto this hydraulic fluid chamber, operation of the cam is transmitted tothe valve, causing the valve to be subjected to opening/closingoperation, whereas, when the hydraulic fluid is drained from thehydraulic fluid chamber, operation of the cam is no longer transmittedto the valve, with the result that the valve is deactivated.

The passage 36 for hydraulic fluid supply/draining in respect of thevalve deactivating mechanism 35 of the first exhaust valve 32 a and thevalve deactivating mechanism 35 of the first intake valve 31 a isprovided with a first control valve 37 and the passage 38 for hydraulicfluid supply/draining in respect of the valve deactivating mechanism 35of the second exhaust valve 32 b and the valve deactivating mechanism 35a of the second intake valve 31 b is provided with a second controlvalve 39, respectively (see FIG. 3).

FIG. 3 shows the layout of the drive and control systems. In thisFigure, an ECU (control unit) 40 for engine control, comprising amicrocomputer etc, inputs signals from an airflow sensor 19 and O₂sensor 23; in addition, it inputs signals from a engine speed sensor 47that detects the engine rotational speed, and an accelerator pedalstroke sensor 48 etc that detects the degree of opening of theaccelerator (amount of depression of the accelerator pedal), forascertaining the operating condition. Also, control signals are outputfrom this ECU 40 to the fuel injection valves 9, the actuators 18 of themultiple linked throttle valves 17 and the first and second controlvalves 39 mentioned above.

The ECU 40 comprises operating condition identifier 41, valve stopmechanism controller 42, air intake rate controller 43 and combustioncontroller 44.

The operating condition identifier 41 ascertains which region theoperating condition is in i.e. whether the operating condition is in theoperating region A on the low load, low rotational speed side, or in theoperating condition B on the high load or high rotational speed side, asshown in FIG. 4, by examining the operating condition of the engine(engine rotational speed and engine load) by means of the signals fromthe engine speed sensor 45 and accelerator pedal stroke sensor 46 etc.Then, based on the results of this determination, in the operatingregion A on. the low load, low rotational speed side, a specialoperating mode is selected in which combustion is effected with theburnt gas exhausted from the preceding (leading) cylinders in theexhaust stroke fed directly to the following cylinders, which are in theintake stroke but, in the operating region B on the high load or highrotational speed side, the ordinary operating mode is selected in whichcombustion is effected with the cylinders respectively operatedindependently.

The valve stop mechanism controller 42 controls the valve deactivationmechanisms 35 as follows by controlling the control valves 37, 39depending on whether the operating condition is in the operating regionA on the low load, low rotational speed side (i.e. the case where thespecial operating mode is selected) or is in the operating region B onthe high load or high rotational speed side (i.e. in the case where theordinary operating mode is selected).

Operating region A: condition in which the first exhaust valve 32 a andfirst intake valve 31 a are deactivated

condition in which the second exhaust valve 32 b and the second intakevalve 31 b are actuated

Operating condition B: condition in which the first exhaust valve 32 aand first intake valve 31 a are actuated

condition in which the second exhaust valve 32 b and the second intakevalve 31 b are deactivated

Flow path changeover means whereby the gas flow path is changed over asdescribed in detail hereinbelow is constituted by this valve stopmechanism controller 42 and the valve deactivation mechanisms 35 whichare controlled thereby.

The air intake rate controller 43 controls the degree of opening (degreeof throttle opening) of the throttle valves 17 by controlling theactuators 18, finds a target air intake rate from a map or the like inaccordance with the operating condition and controls the degree ofthrottle opening in accordance with this target air intake rate. In thiscase, in the low load, low operating speed operating region A, as willbe described, in a condition with introduction of intake air from thebranched intake passage 16 in the following cylinders (second and thirdcylinders 2B and 2C) being cut off, combustion is performed with theratio between the excess air in the gas that is introduced from thepreceding cylinders and the fuel that is newly supplied being a leanair/fuel ratio, the degree of opening of the throttle is adjusted suchthat air that is more by a prescribed amount than the quantity of airthat is necessary for combustion of the fuel in accordance with therequired torque for the two leading and following cylinders (i.e. air ofa quantity in the stoichiometric air/fuel ratio (or theoretical air/fuelratio) with respect to the quantity of fuel for the two cylinders) issupplied to the preceding cylinders (first and fourth cylinders 2A and2D).

The combustion controller 44 comprises fuel injection controller 45 andignition controller 46; the fuel injection controller 45 controls thefuel injection rate from the fuel injection valves 9 provided in eachcylinder 2A to 2D and the injection timing in accordance with theoperating condition of the engine; the injection controller 46 performscontrol such as control of the ignition time and ignition deactivationetc in accordance with the operating condition. Thus, control ofcombustion (control of fuel injection and control of ignition) isaltered in particular in the case where the operating condition is inoperating region A and where the operating condition is in operatingcondition B in FIG. 4.

Specifically, when the operating condition is in operating region A onthe low load, low rotational speed side, the fuel injection rate iscontrolled such that the air/fuel ratio is a lean air/fuel ratio, largerthan the stoichiometric air/fuel ratio, preferably about twice or morethe stoichiometric air/fuel ratio, and, in the compression step, aninjection timing is set such that the fuel injection results in layeringof the mixture and an ignition timing is set such that the forcedignition is performed in the vicinity of the compression top deadcenter. In contrast, in the case of the following cylinders (second andthird cylinders 2B and 2C), fuel is supplied in respect of the burnt gasof the lean air/fuel ratio introduced from the preceding cylinders and,also after supply of fuel, the fuel injection rate is controlled and, inthe intake step, the injection timing is set such that fuel is injectedsuch as to produce an air/fuel ratio that is leaner than thestoichiometric air/fuel ratio and forced ignition is deactivated so thatcompression ignition is performed.

Also, when the operating condition is in the operating region B on thehigh load side or high rotational speed side, the fuel injection rate iscontrolled such that the air/fuel ratio of the cylinders 2A to 2D is ator below the stoichiometric air/fuel ratio: for example, in most of theregion of this operating region B, the air/fuel ratio may be made to bethe stoichiometric air/fuel ratio and, in the fully open load operatingregion and the vicinity thereof, may be made to be richer than thestoichiometric air/fuel ratio. In this case, the injection timing is setsuch that a uniform mixture is produced by injection of fuel to thecylinders 2A to 2D in the intake step and such that forced ignition isproduced in the cylinders 2A to 2D.

The operation of a device according to this embodiment as describedabove will now be described with reference to FIG. 5 to FIG. 7.

In the operating region A on the low load, low rotational speed side, aflow path of substantially new air and gas as shown in FIG. 6 isproduced by putting the device in a special operating mode, in which, asdescribed above, the first exhaust valve 32 a and the first intake valve31 a are in deactivated condition and the second exhaust valve 32 b andsecond intake valve 31 b are in actuated condition. Thus, a two-cylinderconnection condition is produced whereby burnt gas that is exhaustedfrom the preceding cylinders (first and fourth cylinders) 2A to 2D isdirectly introduced into the following cylinders (second and thirdcylinders) 2B and 2C through the inter-cylinder gas passages 22 and onlythe gas exhausted from these following cylinders 2B, 2C is fed to theexhaust passage 20.

In this condition, new air is introduced (arrow a in FIG. 6) from theintake passages 15 in the respective intake strokes to the precedingcylinders 2A, 2D and fuel is injected in the compression stroke whilstperforming feedback control of the fuel injection rate such that theair/fuel ratio detected by a linear O₂ sensor 25 in the precedingcylinders 2A, 2D is a super-lean air/fuel ratio of substantially twiceor more of the stoichiometric air/fuel ratio and stratified chargecombustion is effected (see FIG. 5) with this super-lean air/fuel ratioby performing ignition at a prescribed ignition time.

After this, in the period in which the intake stroke of the precedingcylinders 2A and 2D and the exhaust stroke of the following cylinders 2Band 2C overlap, burnt gas exhausted from the preceding cylinders 2A, 2Dis fed to the following cylinders 2B, 2C through the gas passages 22(white arrow in FIG. 5 and arrow b in FIG. 6). Then, in the followingcylinders 2B, 2C, fuel is supplied to the burnt gas of lean air/fuelratio introduced from the preceding cylinders 2A, 2D and fuel isinjected in the intake step whilst controlling the fuel injection ratesuch as to produce an air/fuel ratio that is leaner than thestoichiometric air/fuel ratio, and compression ignition is thenperformed by rise of pressure and temperature in the combustion chamberin the vicinity of the top dead center of the compression stroke.

In this case, since the high temperature burnt gas that is exhaustedfrom the preceding cylinders 2A, 2D is immediately introduced into thefollowing cylinders 2B, 2C through the short inter-cylinder gas passages22, the temperature in the combustion chamber in the intake stroke inthe following cylinders 2B, 2C becomes high, so that the temperaturewithin the combustion chamber rises to an extent such as to comfortablyenable self-ignition of the mixture in the vicinity of the top deadcenter of the final period of the compression stroke by a further risein pressure and temperature from this condition in the compressionstroke. Furthermore, since the aforesaid burnt gas is thoroughly mixedand uniformly distributed during the period from its being exhaustedfrom the preceding cylinders 2A, 2D up to its being introduced into thefollowing cylinders 2B, 2C, and, in addition, the fuel that is injectedin the intake stroke is uniformly dispersed in the entire combustionchamber during the final period of the compression stroke, a uniformmixture distribution condition that satisfies the ideal simultaneouscompression ignition condition is obtained.

Thus, in the following cylinders 2B, 2C, a large amount of burnt gasconstituents corresponding to EGR gas is contained and, even under thecondition that the air/fuel ratio is lean, combustion is performedrapidly by simultaneous compression ignition; thermal efficiency isthereby greatly increased.

That is, in the preceding cylinders 2A, 2D, thermal efficiency is raisedby stratified charge combustion with a super-lean ratio and pumping lossis lowered; moreover, in the following cylinders 2B, 2C, thermalefficiency is raised by performing compression ignition in a uniformmixture condition while keeping the air/fuel ratio lean and the benefitof lowered pumping loss is obtained in the same way as with thepreceding cylinders 2A, 2D. By these actions, fuel consumption isgreatly improved.

Furthermore, since compression ignition in the following cylinders 2B,2C is achieved by utilizing the heat of the burnt gas that is exhaustedfrom the preceding cylinders 2A, 2D, there is no need to provide specialheating means or to greatly elevate the engine compression ratio andcompression ignition can be effectively performed over a wide operatingrange.

After combustion in the following cylinders 2B, 2C, the gas is exhaustedto the exhaust passage 20 and cleansing of the exhaust gas is performedby a lean NOx catalyst 24A etc provided in the exhaust passage 20.

Also, the rate of generation of NOx in the preceding cylinders 2A, 2D isrestrained to a comparatively low level by making the air/fuel ratiotherein a lean air/fuel,ratio of substantially twice or more thestoichiometric air/fuel ratio and generation of NOx in the followingcylinders 2B, 2C is fully satisfactorily restrained by producing acondition equivalent to that in which a large quantity of EGR isperformed, by introducing burnt gas from the preceding cylinders 2A, 2D.This is also advantageous in reducing emissions.

In contrast, in the operating region B on the high load side or highrotational speed side, the device is put in the ordinary operating mode,in which, as described above, the first exhaust valve 32 a and firstintake valve 31 a are put into actuated condition while the secondexhaust valve 32 b and second intake valve 31 b are put into deactivatedcondition, thereby producing a flow path for substantially new air andgas as shown in FIG. 7, in which the intake ports 31, 31 a and exhaustports 12 a, 12 of the cylinders 2A to 2D are substantially independent,so that new air is introduced into the intake ports 31, 31 a of thecylinders 2A to 2D from the intake passage 15 and burnt gas is exhaustedto the exhaust passage 20 from the exhaust ports 31, 31 a of thecylinders 2A to 2D. In this case, output performance is ensured bycontrolling the intake air rate and fuel-injection rate such that theair/fuel ratio is the stoichiometric air/fuel ratio or is richer thanthis.

In the above embodiment, in the low rotational speed, low load operatingregion A, the fuel injection rate is controlled such that the air/fuelratio of the following cylinders 2B, 2C is leaner than thestoichiometric air/fuel ratio, but it would also be possible to controlthe fuel injection rate such that the air/fuel ratio in the followingcylinders 2B, 2C is at or below or the stoichiometric air/fuel ratio. Inthis case, as shown in FIG. 8, only a three-way catalyst 24B is providedin the exhaust passage 20, or an oxidation catalyst may be provided.Preferably, also, the O₂ sensor 23 provided in the merging section ofthe exhaust passage 20 is a λO₂ sensor whose output changes abruptly inthe vicinity of the stoichiometric air/fuel ratio and the fuel injectionrate to the following cylinders 2B, 2C is subjected to feedback controlin accordance with the output of this O₂ sensor 23. In addition, alinear O₂ sensor 25 whose output changes in linear fashion in responseto the oxygen concentration is provided in the inter-cylinder gaspassage 22 and the fuel injection rate to the preceding cylinders 2A, 2Dwhose air/fuel ratio is made to be a prescribed lean air/fuel ratio issubjected to feedback control in accordance with the output thereof.

In this way, since only gas of the stoichiometric air/fuel ratio isexhausted to the exhaust passage 20 from the following cylinders 2B, 2C,there is no need to provide a lean NOx catalyst and fully satisfactoryexhaust cleansing performance can be secured simply by means of athree-way catalyst 24B (or oxidation catalyst).

Also, since there is no need to provide a lean NOx catalyst, there is noneed to perform temporary enrichment of the air/fuel ratio in order torelease or reduce NOx when the amount of NOx occlusion of the lean NOxcatalyst builds up, so compromise of the improvement in fuel costs isavoided. Furthermore, the problem of sulfur poisoning of the lean NOxcatalyst cannot occur.

Also, although, in the above embodiment, the fuel injection valves ofthe various cylinders were of the direct injection type in which fuel isdirectly injected into the combustion chamber, in regard to thefollowing cylinders it would be possible for fuel to be injected in theintake stroke even in the operating region A of low load, low rotationalspeed, so the fuel injection valves provided in the following cylinderscould be of a type whereby fuel is injected into the intake passagethrough the intake port.

Also, although, in the above embodiment, it was arranged for the flowpath of the new air and gas to be changed over by the flow pathchangeover means in accordance with whether the operating condition wasin the operating region A on the low load, low rotational speed side orwhether the operating condition was in the operating region B on thehigh load side or high rotational speed side, it would also be possibleto arrange for the flow path of new air and gas to be in the aforesaidcondition with two cylinders connected throughout the entire operatingregion.

FIG. 9 shows the setting of operating regions for control in accordancewith a further embodiment of the present invention. In this embodiment,the operating region which was put in two-cylinder connected conditionis divided into a plurality of regions and the air/fuel ratio of thepreceding cylinders (ratio of the fuel injection rate in the case of thepreceding cylinders and the fuel injection rate in the case of thefollowing cylinders) is changed in accordance with these regions.

In this embodiment also, the entire engine is constituted as in FIG. 1or FIG. 8. Also, the control and drive systems are constituted as inFIG. 3 and the operating condition identifier 41 included in the ECU 40ascertains which operating region the operating condition is in i.e.whether the operating condition is in the operating region A on the lowload, low rotational speed side or in the operating condition B on thehigh load or high rotational speed side as shown in FIG. 10. However, inaddition, when in the operating region A in which the special operatingmode is selected, it is arranged to ascertain whether the operatingcondition is in a low speed region A1, intermediate speed region A2 orhigh speed region A3 of this region A.

Also, as the control exercised in the special operating mode, when theoperating condition is in the operating region A on the low load, lowrotational speed side, the combustion condition controller 44 includedin the ECU 40 controls the fuel injection rate in respect of thepreceding cylinders (first and fourth cylinders 2A and 2D) such that theair/fuel ratio is a lean air/fuel ratio greater than the stoichiometricair/fuel ratio and, in the compression stroke, sets the injection timingsuch that layering of the mixture is performed by the fuel injection andsets the ignition timing such that forced ignition is performed in thevicinity of the compression top dead center. On the other hand, inrespect of the following cylinders (second and third cylinders 2B and2C) fuel is supplied in respect of the burnt gas of the lean air/fuelratio that is introduced from the preceding cylinders and the fuelinjection rate is controlled such that the air/fuel ratio issubstantially the stoichiometric air/fuel ratio and also the injectiontiming is set such that fuel is injected in the intake stroke and forcedignition is deactivated in order to perform compression self-ignition.

Furthermore, in this operating region A, the ratio of the fuel injectionrate in respect of the preceding cylinders (first and fourth cylinders2A, 2D) and the fuel injection rate in respect of the followingcylinders (second and third cylinders 2B, 2C) is altered in accordancewith the operating condition such that compression self-ignition issatisfactorily performed in the following cylinders, while adjusting thesum of the fuel injection rates in respect of both of the pair ofcylinders to a rate at which the air/fuel ratio is the stoichiometricair/fuel ratio for the rate of introduction of air to the precedingcylinders.

Specifically, in the intermediate speed region A2 of this operatingregion A, it is arranged to make the air/fuel ratio during combustion inthe preceding cylinders of the order of twice the stoichiometricair/fuel ratio (λ=approximately 2, expressed in terms of the excess airratio λ, when A/F≈30) or more than twice the stoichiometric air/fuelratio (excess air ratio λ is λ>2), by making the fuel injection rate inrespect of the preceding cylinders and the fuel injection rate inrespect of the following cylinders substantially the same, or by makingthe fuel injection rate in respect of the following cylinders a littlelarger. Also, in the low speed region A1 of this operating region A, itis arranged to make the air/fuel ratio during combustion in thepreceding cylinders less than twice the stoichiometric air/fuel ratio(the air excess ratio λ is 1<λ<2), for example A/F≈25, by making thefuel injection rate in respect of the preceding cylinders more than thefuel injection rate in respect of the following cylinders. And in thehigh speed region A3 of this operating region A, it is arranged to makethe air/fuel ratio during combustion in the preceding cylinders lessthan twice the stoichiometric air/fuel ratio (the air excess ratio λ is1<λ<2), for example A/F≈25, by making the fuel injection rate in respectof the preceding cylinders more than the fuel injection rate in respectof the following cylinders.

Next, the operation of a device according to this embodiment will bedescribed.

In a device according to this embodiment also, in the low load, lowrotational speed operating region A, the device is put into the specialoperating mode, in which combustion is effected in the two-cylinderconnected condition, and, in the high load or high rotational speedoperating region, the device is put into the ordinary operating mode, inwhich combustion is performed with the intake ports and exhaust ports ofthe respective cylinders in an independent condition. Thus, in thespecial operating mode, combustion in the preceding cylinders isconducted with a super-lean air/fuel ratio whereas, in the followingcylinders, combustion is conducted by compression self-ignition.

In particular, in the special operating mode, compression self-ignitioncan be performed effectively over a wide operating range, by adjustingthe ratio of the fuel injection rate in respect of the precedingcylinders (first and fourth cylinders 2A and 2D) and the fuel injectionrate in respect of the following cylinders (second and third cylinders2B, 2C) as described above in accordance with the operating condition.

That is, in the low speed region A1 of the operating region A in whichthe special operating mode is conducted, basically a condition obtainsin which the temperature in the combustion chamber is lower than in thecase of the intermediate and high speed regions A2 and A3, socompression self-ignition is difficult to carry out, but, in the lowspeed region A1, control is exercised such that the air/fuel ratio ofthe preceding cylinders is a value of less than twice the stoichiometricair/fuel ratio by making the fuel-injection rate in respect of thepreceding cylinders larger than that of the following cylinders, whileadjusting the air/fuel ratio during combustion in the followingcylinders to be substantially of the stoichiometric air/fuel ratio, so,compared with the case where the air/fuel ratio of the precedingcylinders is made to be twice the stoichiometric air/fuel ratio (i.e.the case where the injection rates of the preceding cylinders andfollowing cylinders are the same), the temperature of the gas that isfed into the following cylinders from the preceding cylinders is raised.As a result, compression self-ignition is performed effectively even inthe low speed region A1.

Also, in the high speed region A3 of the operating region A that hasbeen put in the special operating mode, excessive rise of the combustiontemperature would tend to produce knocking, but, in this region also,the fuel injection rate in respect of the preceding cylinders is madelarger than that in respect of the following cylinders so that theair/fuel ratio of the preceding cylinders is controlled to be a valuesmaller than twice the stoichiometric air/fuel ratio. In this way,although the temperature of the gas that is introduced into thefollowing cylinders rises compared with the case where the air/fuelratio of the preceding cylinders is made to be twice the stoichiometricair/fuel ratio (i.e. the case where the injection rates of the precedingcylinders and following cylinders are the same), the amount of burnt gasconstituents corresponding to EGR in the gas that is introduced into thefollowing cylinders is increased and the fuel injection rate in respectof the following cylinders becomes small. As a result, the energy thatis generated by combustion in the following cylinders becomes small, soknocking is suppressed.

Thus, although, if, by making the fuel injection rate in respect of thepreceding cylinders more than that in respect of the followingcylinders, the air/fuel ratio of the preceding cylinders is controlledso as to be a value smaller than twice the stoichiometric air/fuelratio, this is advantageous in respect of compression self-ignition andknocking prevention compared with the case where the air/fuel ratio ofthe preceding cylinders is made to be twice the stoichiometric air/fuelratio (i.e. the case where the injection rates of the precedingcylinders and following cylinders are the same), on the other hand, itis somewhat disadvantageous in regard to achievement of improved fuelcosts by a stratified charge lean burn in the preceding cylinders and inregard to torque balance between the leading and following cylinders.Accordingly, in the intermediate speed region A2 in which compressionself-ignition of the following cylinders is easily produced by thespecial operating mode and knocking is unlikely, the fuel injection rateis controlled so as to produce a value of the air/fuel ratio of thepreceding cylinders of substantially twice the stoichiometric air/fuelratio, or a value larger than this, so as to be advantageous in respectof improvement of fuel costs and torque balance.

It should be noted that, although, in the example shown in FIG. 9, theoperating region A that was put into the special operating mode wasdivided into a low speed region A1, intermediate speed region A2 andhigh speed region A3 and the air/fuel ratio of the preceding cylinders(i.e. the ratio of the fuel injection rate in respect of the precedingcylinders and the fuel injection rate in respect of the followingcylinders) altered in these regions A1, A2 and A3, it would also bepossible, as shown in FIG. 10, to divide the operating region A that wasput into the special operating mode into a low load region A11,intermediate load region A12 and high load region A13. In this case,control of the fuel injection rate is performed such that in theaforesaid intermediate load region A12 the air/fuel ratio of thepreceding cylinders is made to be a value of substantially twice thestoichiometric air/fuel ratio or a value larger than this, while theair/fuel ratio of the preceding cylinders in the low load region A11 andhigh load region A13 is made to be a value less than twice thestoichiometric air/fuel ratio (for example A/F≈25).

Alternatively, as shown in FIG. 11, in the intermediate load region A20of the operating region A that was put in the special operating mode,the air/fuel ratio of the preceding cylinders may be controlled to avalue that is substantially twice the stoichiometric air/fuel ratio oris larger than this and in the other operating regions the air/fuelratio of the preceding cylinders may be controlled to be a value that issmaller than twice the stoichiometric air/fuel ratio.

In these examples also, in the low load region etc of the operatingregion A that was put in the special operating mode, in which thetemperature in the combustion chamber is comparatively low, compressionself-ignition becomes possible due to the rise in temperature of the gasthat is introduced into the following cylinders from the precedingcylinders and knocking is suppressed due to the reduced generation ofenergy of the following cylinders in the high load region etc whereknocking is likely to occur; a condition is also produced that isbeneficial in respect of improvement of fuel costs and torque balance inthe intermediate load region A12 or intermediate speed, intermediateload region A20.

Although, in the examples shown in FIG. 9, FIG. 10 and FIG. 11 above, ina plurality of operating regions in an operating region A that was putinto the special operating mode, it was arranged to change over theair/fuel ratio of the preceding cylinders between a value ofsubstantially twice the stoichiometric air/fuel ratio or larger thanthis and a value smaller than this, it could be arranged to change theair/fuel ratio of the preceding cylinders progressively in accordancewith the operating condition whilst keeping the air/fuel ratio largerthan the stoichiometric air/fuel ratio.

In this case, in at least a low load region of the operating region A,the air/fuel ratio during combustion in the preceding cylinders is madesmaller as the load becomes lower. Alternatively, in at least a lowspeed region of the operating region A, the air/fuel ratio duringcombustion in the preceding cylinders is made smaller as the speedbecomes lower.

For example, when the likelihood of knocking on the high speed, highload side of the operating region A which has been put into the specialoperating mode has been reduced by the provision of cooling means in theinter-cylinder gas passages 22, as shown in FIG. 12, the air/fuel ratioof the preceding cylinders may be made to be a value of substantiallytwice the stoichiometric air/fuel ratio or a value larger than this onthe high speed, high load side of the operating region A which has beenput into the special operating mode and the air/fuel ratio of thepreceding cylinders may be arranged to be changed to the rich side andas the engine rotational speed or load becomes lower.

In this way, in the operating region A that was put into the specialoperating mode, the condition in which compression self-ignition ispossible may be ensured by raising the temperature of the gas that isfed into the following cylinders from the preceding cylinders so as tocompensate for the tendency for the temperature within the combustionchamber of the following cylinders to become lower as the enginerotational speed (or load) becomes lower.

Also, as shown in FIG. 13, control may be exercised such as to make theair/fuel ratio of the preceding cylinders in the intermediate speed,intermediate load region A20 of the operating region A that was put inthe special operating mode a value of substantially twice thestoichiometric air/fuel ratio or greater than this and to progressivelyreduce the air/fuel ratio during combustion in the preceding cylindersas this region is departed from towards the low speed, low load side(direction of the arrow a) or towards the high speed, high load side(direction of the arrow b).

In this way, on the low speed, low load side of the operating region Athat is put into the special operating mode, an excellent effect interms of ensuring a condition in which compression self-ignition ispossible and an excellent effect in terms of suppressing knocking on thehigh speed, high load side are obtained.

Also, in addition to control in accordance with the operating conditionin the operating region A that was put in the special operating mode asdescribed above, it may be arranged to alter the air/fuel ratio of thepreceding cylinders in accordance with the temperature condition of theengine. For example, in cases where the engine temperature is low evenafter engine warm-up (cases where the temperature of the engine coolingwater is below the prescribed temperature) it is preferable to make theair/fuel ratio of the preceding cylinders less than twice thestoichiometric air/fuel ratio in the entire region in the operatingregion A that has been put into the special operating mode. In this way,it is possible to ensure a condition in which compression self-ignitionis possible by raising the temperature of the gas that is introducedinto the following cylinders from the preceding cylinders even when theengine temperature is comparatively low.

Also, although, in the above examples, it is arranged to performcombustion by compression self-ignition in the following cylinders overthe entire region of the operating region A that was put into thespecial operating mode, it would also be possible to arrange to performcombustion in the following cylinders by forced ignition by performingignition using a spark plug 7 in a prescribed ignition period in part ofthe operating region A that was put into the special operating mode, forexample in an extremely low speed, low load region where it is difficultfor the temperature and pressure within the combustion chamber toachieve a condition in which compression self-ignition is possible.Alternatively, it would also be possible to arrange to performcombustion by forced ignition in the following cylinders when the enginetemperature is low.

FIG. 14 to FIG. 16 show control of intake/exhaust and combustion inaccordance with the operating condition according to a furtherembodiment of the present invention.

In this embodiment also, the engine as a whole is constituted as shownin FIG. 1 or FIG. 8. Also, the control/drive system is constituted as inFIG. 3 and the operating condition identifier 41 included in the ECU 40ascertains whether the operating condition is in the operating region Aon the low load, low rotational speed side as shown in FIG. 14 is in thehigh load or high rotational speed side operating region B. However, inaddition, when the operating condition is in the partial load region Ain which the special operating mode is selected, a function is providedof ascertaining whether it is in the high load side region A102 of thisregion A or the region other than this i.e. the low load side regionA101 of this partial load region A.

Also, when the operating condition is in the operating region A on thelow load, low rotational speed side, the combustion condition controller44 included in the ECU 40 controls the fuel injection rate in respect ofthe preceding cylinders (first and fourth cylinders 2A, 2D), byexercising control in the special operating mode, such as to make theair/fuel ratio a lean air/fuel ratio larger than the stoichiometricair/fuel ratio and sets the injection timing in the compression strokesuch that the layering of the mixture is produced by the fuel injectionand sets the ignition timing such that forced ignition is performed inthe vicinity of the compression top dead center. On the other hand, inrespect of the following cylinders (second and third cylinders 2B, 2C),it supplies fuel with respect to the burnt gas of lean air/fuel ratiointroduced from the preceding cylinders and controls the fuel injectionrate such that the air/fuel ratio is substantially the stoichiometricratio and sets the injection timing such that fuel is injected in theintake stroke and deactivates forced ignition so that compressionself-ignition is performed.

Also, in the low load side region A101 of the operating region A inwhich the control of the aforesaid special operating mode is executed,as shown in FIG. 15 by the solid line, the injection timing is set suchthat fuel is injected in the intake stroke of the following cylinders2B, 2C and, in the high load side region A2 of the operating region A,the injection time of the fuel with respect to the following cylinders2B, 2C is set to the latter half of the compression stroke of thefollowing cylinders 2B, 2C i.e. to a time close to the compression topdead center PTDC, as shown by the broken line in FIG. 15, by relativelyretarding the injection time of the fuel compared with the aforesaid lowload side region A1. In FIG. 5, the times indicated by the symbols T31,T32 b, T31 b and T32 respectively indicate the valve-opening periods ofthe intake valve 31, second exhaust valve 32 b, second intake valve 31 band exhaust valve 32.

Next, the operation of this embodiment will be described.

In the device of this embodiment also, in the low load, low rotationalspeed operating region A, the device is put into the special operatingmode in which combustion is effected in the two-cylinder connectedcondition and, in the operating condition on the high load side or highrotational speed side, the device is put into the ordinary operatingmode in which combustion is effected with the intake ports and exhaustports of the respective cylinders put in an independent condition. Also,when in the special operating mode, combustion is performed in thepreceding cylinders with a super-lean air/fuel ratio whereas in thefollowing cylinders combustion is performed by compressionself-ignition.

Also, in the operating region A in which control is performed inaccordance with the special operating mode as described above, byadjusting the fuel ignition time in respect of the following cylinders2B, 2C as described above in accordance with the operating condition, itis possible to ensure that compression self-ignition can be effectivelyperformed over a wide operating range without occurrence of knocking.

Specifically, although, in the low speed side region A101 of theoperating region A that was put in the special operating mode asdescribed above, the conditions are such that compression self-ignitionis more difficult to achieve than in the case of the high load sideregion A102, by setting the ignition time of the fuel at some pointduring the intake stroke as described above, this fuel and air (burntgas of lean air/fuel ratio introduced from the preceding cylinders 2A,2D) are thoroughly mixed so combustibility is promoted, with the resultthat compression self-ignition can be effectively performed even in thelow speed region A101.

Also, although, in the high load side region A102 of the operatingregion A that was put in the aforesaid special operating mode,compression self-ignition occurs easily due to the higher temperature ofthe combustion chamber than in the low load side region A101, on theother hand, knocking of the following cylinders 2B, 2C tends to occur,so, by retarding the injection time of the fuel as described above in aperiod close to the compression top dead center PTDC, compressionself-ignition is not performed until the mixture has been thoroughlyactivated after fuel injection, so that occurrence of knocking, in whichself-ignition of the mixture occurs prior to propagation of the flamethrough the interior of the combustion chamber, is prevented.

Thus, in the operating region A in which compression self-ignition isperformed in the following cylinders 2B, 2C, if it is found that anoperating condition obtains in which knocking can easily occur i.e. theoperating condition of the high load side operating region A102 asdescribed above, it is arranged to suppress activation of the mixture byrelatively retarding the injection time of the fuel with respect to thefollowing cylinders 2B, 2C, so knocking, which occurs due to excessiveease of ignition of the mixture, can be effectively prevented.Furthermore, in the compression self-ignition region of the followingcylinders 2B, 2C, if it is found that an operating condition obtains inwhich knocking is unlikely i.e. that the operating condition of the lowload side region A101 described above obtains, activation of the mixtureis promoted by relatively advancing the ignition time of the fuel withrespect to the following cylinders 2B, 2C, so misfiring in the followingcylinders 2B, 2C is effectively prevented and combustion by compressionself-ignition can be reliably performed. In this way, the benefits areobtained that the thermal efficiency of the engine is improved and thatengine output can be fully guaranteed.

In particular, when an operating condition obtains in which knocking isliable to occur in the compression self-ignition region of the followingcylinders 2B, 2C, as shown in the above embodiment, if the injectiontime of the fuel with respect to the following cylinders 2B, 2C is setin the latter half of the compression stroke, activation of the mixtureis effectively suppressed, thereby making it possible to reliablysuppress occurrence of knocking as described above.

It should be noted that if it is found that an operating conditionobtains in which knocking is liable to occur i.e. the operatingcondition of the high load side operating region A102 in the operatingregion A in which compression self-ignition of the following cylinders2B, 2C is performed, as shown in FIG. 16, it may be arranged to injectthe fuel in divided fashion into the following cylinders 2B, 2C and toset the latter injection time S2 of the fuel in this divided injectiontime in the latter half of the compression. By adopting such anarrangement, the advantage is thereby obtained that occurrence ofknocking can be effectively prevented by suppressing the mixing of fuelcorresponding to the injection of the latter period injected in thelatter injection period F2 with air to a suitable degree, whilemaintaining combustibility by thoroughly mixing the fuel correspondingto the former injection period F1 of the aforementioned dividedinjection time i.e. the former injection period that was injected duringthe course of the intake stroke of the following cylinders 2B, 2C.

Also, it may be arranged that, in the operating region A of thefollowing cylinders 2B, 2C where compression self-ignition is performed,the probability of occurrence of knocking is ascertained in accordancewith the engine load etc and the latter injection period F2 in theaforementioned divided injection time of the fuel is retarded so as toapproach more closely to the compression top dead center as theprobability of occurrence of such knocking becomes higher. If it isarranged in this way for the latter injection period F2 to be changed inaccordance with the probability of occurrence of knocking, it ispossible to effectively prevent occurrence of knocking on the high loadside of the engine, where the temperature in the combustion chambertends to be higher, while effectively preventing occurrence of misfiringon the low load side of the engine, where the temperature in thecombustion chamber tends to be lower.

Furthermore, it may be arranged that, in the operating region A wherecompression self-ignition is performed in the following cylinders 2B,2C, if an operating condition obtains in which knocking is likely tooccur, the fuel is injected into the following cylinders 2B, 2C individed fashion and the amount of the latter injection period of thefuel in this divided injection is set to a value greater than the amountof injection in the former period thereof. This arrangement has theadvantage that if, in the region where compression self-ignition isperformed in the following cylinders, it is found that the operatingcondition is such that knocking is liable to occur due for example to ahigh temperature within the combustion chambers of the followingcylinders 2B, 2C, activation of the mixture in the operating region A2where the probability of knocking is high is effectively suppressed bysetting the amount of injection of fuel in the latter period of thedivided injection of the following cylinders 2B, 2C to a greater valuethan the amount of injection of fuel in the former period thereof, sooccurrence of knocking can thereby be more reliably prevented.

Also, in the operating region A in which compression self-ignition isperformed in the following cylinders 2B, 2C, preferably the probabilityof occurrence of knocking is ascertained and the ratio of the latterinjection period amount with respect to the total injection mount of thefuel injected into the following cylinders is changed so as to increaseas the probability of occurrence of such knocking becomes higher. Withsuch a construction, in the compression self-ignition region of thefollowing cylinders 2B, 2C, if it is arranged for the latter injectionperiod amount of fuel to be changed in accordance with the probabilityof occurrence of knocking, the advantage is obtained that occurrence ofknocking can be effectively prevented on the high load side of theengine, where the temperature in the combustion chamber tends to becomehigh, while also effectively preventing occurrence of misfiring on thelow load side of the engine, where the temperature in the combustionchamber tends to become low.

Also, in the above embodiment, in the operating region A whereincompression self-ignition is performed in the following cylinders 2B,2C, the construction is such that, if the engine is in the high loadoperating region A2, it is determined that the engine is in a conditionin which knocking is likely to occur, so it can be ascertained easilyand appropriately whether or not the temperature in the combustionchambers of the following cylinders 2B, 2C is tending to become high, inaccordance with the engine load. It can therefore be accuratelyascertained from the engine load whether or not the engine is in anoperating condition in which knocking is likely to occur in thefollowing cylinders 2B, 2C and the fuel injection time in respect of thefollowing cylinders 2B, 2C can be controlled appropriately in accordancewith the result of this determination.

It should be noted that determination means to ascertain the octanevalue of the fuel employed could be provided so that it can beascertained in accordance with the result of the determination by thisdetermination means whether the engine is in a condition in whichknocking is likely to occur in the compression self-ignition region ofthe following cylinders 2B, 2C. That is, since knocking is more likelyto occur when the octane value of the fuel employed is lower, it may beconcluded that the engine is in an operating condition in which knockingis likely to occur in the compression self-ignition region of thefollowing cylinders 2B, 2C if this determination means ascertains thatfuel of low octane value is being employed; the fuel-injection time inrespect of the following cylinders 2B, 2C may then be appropriatelycontrolled in accordance with the result of this determination.

Also, in the operating region A in which compression self-ignition isbeing performed in the following cylinders 2B, 2C, if the engine is inan operating condition in which knocking is likely to occur, it isdesirable to adopt a construction in which swirl generating means isprovided that generates swirling such that a high intensity ofturbulence is maintained in the latter half of the compression stroke.With such a construction, in the compression self-ignition region of thefollowing cylinders 2B, 2C, if the engine is in an operating conditionin which knocking is likely to occur, amelioration of the drop incombustibility caused by retarding of the injection time of the fuel canbe achieved by the swirling that is generated by this swirl generatingmeans. There is therefore the advantage that the benefit of improvementof combustibility produced by maintaining a high intensity of turbulencein the latter half of the compression stroke due to this swirling andthe benefit of suppression of knocking due to the fuel injection time inrespect of the following cylinders being retarded so as to approach moreclosely to the compression top dead center etc can be achieved at sametime.

For example, as shown in FIG. 17, the leading end portion of theinter-cylinder gas passage 22 i.e. the downstream side of theinter-cylinder gas passage 22 that is connected with the second intakeport 11 b of the following cylinders 2B, 2C is arranged so as to bedirected in the cylinder tangential direction of the following cylinders2B, 2C in plan view. Thus, in the intake stroke of the followingcylinders 2B, 2C, the burnt gas of the preceding cylinders 2A, 2D isintroduced into the inter-cylinder gas passages 22 by opening of thesecond exhaust port 12 b of the preceding cylinders 2A, 2D and burnt gasis introduced into the combustion chambers of the following cylinders2B, 2C along the tangential directions thereof (direction of the arrow bin FIG. 17) from the aforementioned inter-cylinder gas passage 22 byopening of the second intake port 11 b of the following cylinders 2B,2C, thereby enabling swirl to be generated in the combustion chambers ofthe following cylinders 2B, 2C and maintaining the intensity ofturbulence of this swirl at a high level in the latter half of thecompression stroke. Combustibility in the following cylinders 2B, 2C canthereby be effectively improved.

FIG. 18 and FIG. 19 show intake, exhaust and combustion control inaccordance with operating condition in yet another embodiment of thepresent invention.

In this embodiment also, the engine as a whole is constructed as shownin FIG. 1 or FIG. 8. Also, the control/drive system is constructed asshown in FIG. 3 in which the operation condition determination means 41included in the ECU 40 ascertains which region the operating conditionis in, of the operating region A on the low load, low rotational speedside shown in FIG. 18 (operating region in which the special operatingmode is selected) and the operating region B on the high load or highrotational speed side (operating condition in which the ordinaryoperating mode is selected). However, in addition, when the engine is inthe operating region A in which the special operating mode is selected,it is arranged to ascertain, of this region A, whether the engine is ina high load side region A202 or a region A201 on the low load sidethereof.

Also, the combustion condition controller 44 included in the ECU 40,when the operating condition is in the operating region A on the lowload, low rotational speed side, as the control in the special operatingmode, controls the fuel injection rate such that, in regard to thepreceding cylinders (first and fourth cylinders 2A, 2D), the air/fuelratio is made to be a lean air/fuel ratio that is larger than thestoichiometric air/fuel ratio and, in the compression stroke, sets aninjection timing such that layering of the mixture is produced by fuelinjection and sets an ignition timing such that forced ignition isperformed in the vicinity of the compression top dead center. On theother hand, in regard to the following cylinders (second and thirdcylinders 2B, 2C), fuel is supplied with respect to the burnt gas oflean air/fuel ratio that is introduced from the preceding cylinders andthe fuel injection rate is controlled such that substantially thestoichiometric air/fuel ratio is produced and fuel is injected in theintake stroke and forced ignition is deactivated so that compressionself-ignition is performed.

Furthermore, in the above operating region A, the sum of the fuelinjection rates in respect of pairs of cylinders comprising a precedingcylinder and a following cylinder is adjusted to a rate such as toproduce the stoichiometric air/fuel ratio in respect of the rate of newair introduction into the preceding cylinder and the ratio of the fuelinjection rate in respect of the preceding cylinders (first and fourthcylinders) 2A and 2D and the fuel injection rate in respect of thefollowing cylinders (second and third cylinders) 2B and 2C is controlledin accordance with the operating condition such that compressionself-ignition is fully satisfactorily performed, while preventingoccurrence of knocking in the following cylinders.

Specifically, in the low load side region A201 of the operating regionA, by making the fuel injection rate in respect of the precedingcylinders 2A, 2D and the fuel injection rate in respect of the followingcylinders 2B, 2C substantially the same or by making the fuel injectionrate of the following cylinders 2B, 2C a little larger, the air/fuelratio during combustion in the preceding cylinders 2A, 2D becomes abouttwice the stoichiometric air/fuel ratio (A/F≈30, i.e. λ=about 2expressed in terms of air excess ratio λ) or more than twice thestoichiometric air/fuel ratio (air excess ratio λ is λ>2). As a result,in the region A201 on the aforesaid low load side in which misfiring ofthe following cylinders 2B, 2C tends to be liable to occur due to thetotal fuel injection rate being set to a relatively low value due to thelow engine load, setting of the fuel injection rate in respect of thefollowing cylinders 2B, 2C to an excessively low value is prevented andoccurrence of the aforesaid misfiring is thereby prevented.

In this regard, in the high load side region A202 of the aforesaidoperating region A, control is exercised such that the air/fuel ratioduring combustion in the preceding cylinders is less than twice thestoichiometric air/fuel ratio (the air excess ratio λ is 1<λ<2), forexample such that A/F≈25, by making the fuel injection rate in respectof the preceding cylinders 2A, 2D more than the fuel injection rate inrespect of the following cylinders 2B, 2C; the air/fuel ratio of thepreceding cylinders 2A, 2D is thereby made relatively richer than in theregion A1 on the low load side. As a result, in the aforesaid regionA202 on the high load side, where the temperature of the followingcylinders 2B, 2C becomes excessively high due to the total fuelinjection rate being set to a relatively high level due to the highengine load and, corresponding to this, knocking tends to be liable tooccur in the following cylinders 2B, 2C, occurrence of the aforesaidknocking is prevented by the EGR effect by introduction of a largeamount of burnt gas to the following cylinders 2B, 2C.

Also, in the high load side region A2 of the operating region A, asdescribed above, if the fuel injection rate in respect of the precedingcylinders 2A, 2D is set larger than the fuel injection rate in respectof the following cylinders 2B, 2C, there is a risk that implementationof the aforementioned special operating mode control may becomeimpossible, due to its becoming impossible to burn the fuel injectedinto the following cylinders 2B, 2C due to lowering of the oxygenconcentration in the burnt gas introduced into the following cylinders2B, 2C. Consequently, in the high load side region A2 of the aforesaidoperating region A, control is exercised so as to introduce new air intothe following cylinders 2B, 2C in addition to the burnt gas that isintroduced from the preceding cylinders 2A, 2D, by temporary opening ofa new air introduction intake valve (first intake valve 31 a) forintroducing new air into the following cylinders 2B, 2C.

That is, in the high load side region A2 of the aforesaid operatingregion A, after opening of the first intake valve 31 a in the vicinityof the intake top dead center of the following cylinders 2B, 2C, thisfirst intake valve 31 a is put into the closed condition during thecourse of the subsequent intake stroke of the following cylinders 2B,2C. Also, the burnt gas introduction valve (second intake valve 31 b) ofthe following cylinders 2B, 2C is maintained in closed condition untilimmediately prior to the putting of the first intake valve 31 a into theclosed condition; only then is the burnt gas that was introduced fromthe preceding cylinders 2A, D introduced into the following cylinders2B, 2C by opening of the burnt gas introduction valve.

Next, the operation of a device according to this embodiment will bedescribed.

In the device according to this embodiment also, in the operating regionA of low load, low rotational speed, the special operating mode isproduced by performing combustion in the two-cylinder connectedcondition and, in the operating region on the high load side or highrotational speed side, ordinary operating mode is produced by conductingcombustion in a condition with the intake ports and exhaust ports of thevarious cylinders made independent. Thus, when in the special operatingmode, combustion is conducted in the preceding cylinders with asuper-lean air/fuel ratio whereas in the following cylinders combustionis conducted by compression self-ignition.

In particular, compression self-ignition can be appropriately conductedover a wide operating region by adjusting, as described above, the ratioof the rate of fuel injection in respect of the preceding cylinders(first and fourth cylinders 2A, 2D) and the rate of fuel injection inrespect of the following cylinders (second and third cylinders 2B, 2C)in the special operating mode in accordance with the operatingcondition.

That is, in the region A202 on the high load side of the operatingregion A that was put into the special operating mode, control isexercised such that the air/fuel ratio of the preceding cylinders 2A, 2Dis made relatively rich i.e. such that it becomes a value of less thantwice the stoichiometric air/fuel ratio, by setting the fuel injectionrate in respect of the preceding cylinders 2A, 2D to be more than in thecase of the region A201 on the low load side. In this way, knocking issuppressed by the EGR effect by increase of the amount of burnt gasconstituents corresponding to EGR in the gas that is introduced into thefollowing cylinders 2B, 2C, albeit the temperature of the gas that isintroduced into the following cylinders 2B, 2C is raised compared withthe case where the air/fuel ratio of the preceding cylinders 2A, 2D ismade to be twice the stoichiometric air/fuel ratio (i.e. the case wherethe injection rates of the preceding cylinders and following cylindersare the same).

Thus, although the amount of new air in the burnt gas that is introducedinto the following cylinders 2B, 2C is reduced by setting the air/fuelratio of the preceding cylinders 2A, 2D to a value smaller than twicethe stoichiometric air/fuel ratio in the aforesaid high load side regionA202, in this case, thanks to the adoption of a construction in whichnew air is introduced into the following cylinders 2B, 2C in addition tothe burnt gas introduced from the preceding cylinders 2A, 2D, thedeficiency of new air in the following cylinders 2B, 2C in the high loadside region A2 is eliminated, so compression self-ignition can beproperly performed.

Specifically, as shown in FIG. 19, the amount of new air needed toperform compression self-ignition in the following cylinders 2B, 2C isensured by adopting a construction in which new air introduced throughthe intake passage 15 and branch passage 16 is supplied to the followingcylinders 2B, 2C by putting the new air introduction intake valve (firstintake valve 31 a) into an open condition whilst maintaining the burntgas introduction valve (second intake valve 31 b) of the followingcylinders 2B, 2C in a closed condition in the vicinity of the intake topdead center (ITDC) of the following cylinders 2B, 2C. The burnt gas thatis introduced from the preceding cylinders 2A, 2D can then be introducedinto the following cylinders 2B, 2C by putting the first intake valve 31a in a closed condition during the intake stroke of the followingcylinders 2B, 2C and putting the second intake valve 31 b of thefollowing cylinders 2B, 2C into an open condition prior to this.

As described above, in the high load side region A202 in the compressionself-ignition region A, new air of comparatively low temperature can beefficiently introduced into the following cylinders 2B, 2C prior tointroduction into the following cylinders 2B, 2C of the burnt gasextracted from the preceding cylinders 2A, 2D, by putting the new airintroduction intake valve (first intake valve 31 a) into an opencondition in the vicinity of the intake top dead center (ITDC) of thefollowing cylinders 2B, 2C. Furthermore, the air/fuel ratio of thefollowing cylinders 2B, 2C can be prevented from becoming lean due tothe introduction of new air into the following cylinders 2B, 2C in thelow load region A201 in the compression self-ignition region A of thefollowing cylinders 2B, 2C in which the oxygen concentration in theburnt gas introduced into the following cylinders 2B, 2C was maintainedat a sufficiently high level by making the air/fuel ratio of thepreceding cylinders 2A, 2D comparatively lean, by arranging that in thelow load side region A201 in the compression self-ignition region(partial load region) A of the following cylinders 2B, 2C the new airintroduction intake valve (first intake valve 31 a) is maintained in anopen condition.

Also, burnt gas extracted from the preceding cylinders 2A, 2C bystopping introduction of this new air after efficiently introducing newair into the following cylinders 2B, 2C by putting the aforesaid new airintroduction intake valve (first intake valve 31 a), which was opened inthe high load side region A202 in the compression self-ignition region Aof the following cylinders 2B, 2C, into the closed condition during theintake stroke of the following cylinders 2B, 2C can be smoothlyintroduced into the following cylinders 2B, 2C.

Furthermore, if, as shown in the above embodiment, in the high load sideregion A202 of the compression self-ignition region A of the followingcylinders 2B, 2C, it is arranged to open the burnt gas introductionvalve (second intake valve 31 b) of the following cylinders 2B, 2Cduring the intake stroke and to open the new air introduction intakevalve (first intake valve 31 a) prior to the valve opening time of thisburnt gas introduction valve (second intake valve 31 b), for example inthe vicinity of the intake top dead center (ITDC) of the followingcylinders 2B, 2C, in the high load side region A202 in the aforesaidpressure self-ignition region A, new air can be efficiently introducedinto the following cylinders 2B, 2C and the burnt gas introduced fromthe preceding cylinders 2A, 2D can be efficiently introduced into thefollowing cylinders 2B, 2C by putting the aforesaid new air introductionintake valve (first intake valve 31 a) into the closed condition.

Specifically, although, as shown by the broken line in FIG. 19, it wouldbe possible to put the burnt gas introduction valve (second intake valve31 b) into the open condition in the vicinity of the intake top deadcenter ITDC of the following cylinders 2B, 2C, if this were done, therate of introduction of new air would be decreased due to the new airsupplied from the intake passage 15 and the burnt gas supplied throughthe inter-cylinder gas passage 22 into the following cylinders 2B, 2Cbeing introduced simultaneously. It is therefore preferable to arrangethat the new air should be efficiently introduced into the followingcylinders 2B, 2C by maintaining the burn gas introduction valve (secondintake valve 31 b) in closed condition up to a point during the intakestroke of the following cylinders 2B, 2C, as described above. Also, ifit is arranged that the burnt gas introduction valve is maintained inclosed condition up to a point during the intake stroke of the followingcylinders 2B, 2C, there is the advantage that compression self-ignitioncan be achieved by a rise of the internal temperature of the precedingcylinders 2A, 2D due to increase in the amount of internal EGR in thepreceding cylinders 2A, 2D.

Also, if it is arranged that, in the high load side region A202 in thecompression self-ignition region A of the following cylinders 2A, 2D,control is exercised so as to raise the ratio of the rate ofintroduction of new air with respect to the total rate of introductionof gas into the following cylinders 2B, 2C in response to enrichment ofthe air/gas ratio of the preceding cylinders 2A, 2D, by more than thiscompared with the low load side region A201, the advantage is obtainedthat, if the oxygen concentration in the burnt gas that is introducedinto the following cylinders 2B, 2C drops in response to the air/fuelratio of the preceding cylinders 2A, 2D being set to be comparativelyrich in the high load side region A201 in the compression self-ignitionregion A of the following cylinders 2B, 2C, the engine output can bemaintained at a fully satisfactory level and the occurrence of knockingcan be effectively prevented by suppressing the rise in temperature inthe following cylinders 2B, 2C, by appropriately conducting compressionself-ignition of the following cylinders 2B, 2C by effectivelyeliminating insufficiency of new air in the following cylinders 2B, 2C,thanks to this raising of the ratio of the rate of introduction of newair with respect to the total gas introduction rate into the followingcylinders 2B, 2C,

Also, if it is arranged that the air/fuel ratio of the followingcylinders 2B, 2C is controlled such that the oxygen concentration in theexhaust gas that is exhausted from the following cylinders 2B, 2C atleast in the compression self-ignition region A of the followingcylinders 2B, 2C is a value corresponding to the combustion condition ofthe stoichiometric air/fuel ratio, only burnt gas of the followingcylinders 2B, 2C that was burnt with the stoichiometric air/fuel ratiowhile combustion in the preceding cylinders 2A, 2D is being conductedwith a lean air/fuel ratio is fed out to the exhaust passage 20. Thereis therefore no need to provide a lean NOx catalyst as in a conventionallean-burn engine, so exhaust cleansing performance can be fullysatisfactorily ensured simply by a three-way catalyst 24. Thus, sincethere is no need to provide a lean MOX catalyst, there is no need fortemporary enrichment of the air/fuel ratio for purposes of release andreduction of NOx when the amount of NOx occluded by the lean NOxcatalyst builds up, so compromise of the improvement in fuel costs isavoided. Furthermore, the problem of sulfur poisoning of the lean NOxcatalyst cannot occur.

FIG. 20 shows a drive/control system according to yet a furtherembodiment of the present invention. In this embodiment, the ECU 50comprises as its functional elements operating condition identifier 51,temperature status identifier 52, mode setting means 53, valve stopmechanism controller 54, intake air quantity controller 55 andcombustion controller 56.

The operating condition identifier 51 determines whether the operatingcondition is in the operating region A on the low load, low rotationalspeed side as shown in FIG. 21 or in the operating region B on the highload or high rotational speed side and furthermore, if the operatingcondition is in the special operating mode region A, determines whetherit is in the low load side operating region A301, the intermediate loadside operating region A302 or the high load side operating region A303of this region A.

The temperature condition determining means 52 ascertains thetemperature condition of the engine by means of a signal from a watertemperature sensor 49 and ascertains whether or not compressionself-ignition in the following cylinders will be difficult, based onthis engine temperature and in particular on the temperature of thefollowing cylinders. Specifically, the temperature status identifier 52ascertains whether the water temperature (engine temperature) is lowtemperature below a prescribed value or is high temperature higher thanthe prescribed temperature. It should be noted that this temperaturestatus identifier 52 is not restricted to ascertaining the temperaturecondition of the engine by means of a signal from the water temperaturesensor 49 but could, apart from this, ascertain the temperaturecondition of the engine directly or indirectly or could ascertain thetemperature condition of the engine by means of the exhaust gasexhausted from the cylinders, for example by the provision of an exhaustgas temperature sensor.

Based on the determination by the operating condition identifier 51, inthe aforesaid special operating mode region A, the mode selection means53 selects the special operating mode in which combustion is effected byintroducing burnt gas exhausted from the preceding cylinders which arein the exhaust stroke directly into the following cylinders which are inthe intake stroke and, in the aforesaid ordinary operating mode regionB, selects the ordinary operating mode, in which combustion is performedindependently in the respective cylinders.

Also, the mode setting means 53 performs setting such as to change overthe combustion condition between the compression self-ignition mode andforced ignition mode in respect of the following cylinders 2B, 2C andperforms setting so as to change over the combustion condition betweenthe stratified charge combustion mode and uniform lean combustion modein respect of the preceding cylinders 2A, 2D.

Specifically, when the mode setting means 53 selects the specialoperating mode and the operating condition identifier 51 ascertains thatthe engine operating condition is in the low load side operating regionA301, at low temperature the forced ignition mode, in which combustionin these following cylinders 2B, 2C is effected by forced ignition, isselected, for the reason that, in accordance with the determination ofthe engine temperature condition by the temperature status identifier52, it appears that compression self-ignition in the following cylinders2B, 2C will be difficult; and, at high temperature, the compressionself-ignition mode, in which combustion in the following cylinders iseffected by compression self-ignition, is selected, for the reason thatit appears that compression self-ignition in the following cylinderswill be possible. That is, if for example the engine has not been fullywarmed-up, with the result that the temperature of the combustionchambers of the following cylinders 2B, 2C is low, if, even in suchcircumstances, combustion by compression self-ignition were continued inthe following cylinders 2B, 2C, there would be a risk of being unable toguarantee stable combustion, due to misfiring etc. Consequently, in suchcases, forced ignition mode is selected as described above, to ensurethat a stable combustion can be achieved.

Furthermore, if the mode setting means 53 ascertains, by means of theoperating condition identifier 51, that the engine operating conditionis in the intermediate or low load side operating region A301 or A302 ofthe special mode region A, it selects the stratified charge leancombustion mode in which the combustion in the preceding cylinders 2A,2D is put into a stratified lean condition; when in the operating regionA303, in which the engine load is on the high load side compared withthe operating condition where this stratified charge lean combustionmode is selected, it selects the uniform lean combustion mode, in whichthe combustion in the preceding cylinders 2A, 2D is put into the uniformlean condition. Also, even in the intermediate or low load sideoperating regions A301, A302 in which the aforesaid stratified chargelean combustion mode is adopted, if the forced ignition mode isselected, a shift is effected to the uniform lean combustion mode. This“stratified charge lean combustion” means a combustion mode in whichcombustion is effected of a lean mixture with the injected fuel in astratified form and “uniform lean combustion” means a combustion mode inwhich combustion is effected of a lean mixture with the injected fueluniformly dispersed. Thus, control in which a changeover of thecombustion mode in the preceding cylinders 2A, 2D is effected betweenstratified charge lean combustion and uniform lean combustion inaccordance with the load region of the engine i.e. a changeover betweenstratified charge lean combustion mode and uniform lean combustion modeis effected is based on the following characteristics in each combustionmode.

FIG. 22 shows the relationship between burnt gas temperature andair/fuel ratio under the same load in stratified charge lean combustionand uniform lean combustion. From this FIG. 22, it can be seen that,regarding the temperature of the burnt gas at the same air/fuel ratio,that of stratified charge lean combustion is higher than that of uniformlean combustion. Consequently, when high temperature burnt gas is to beintroduced into the following cylinders 2B, 2C, the combustion mode inthe preceding cylinders 2A, 2D is appropriately stratified charge leancombustion; contrariwise, when it is not desired to raise thetemperature of the following cylinders 2B, 2C, the combustion mode inthe preceding cylinders 2A, 2D should appropriately be uniform leancombustion. Also, since the burnt gas temperatures are thus differentfor the same air/fuel ratio, whereas uniform lean combustion hasexcellent thermal efficiency compared with stratified charge leancombustion and so exhibits excellent fuel cost characteristics, on theother hand, with uniform lean combustion, ignition becomes difficult asthe air/fuel ratio becomes larger i.e. as a super-lean condition isapproached, so there are limits to the extent to which the air/fuelratio can be increased. Consequently, in order to improve fuel costs,uniform lean combustion, whose fuel costs characteristic is excellent,is appropriate in the range of air/fuel ratio where uniform leancombustion is feasible but, outside this range, stratified charge leancombustion, in which a super-lean air/fuel ratio can be set, isappropriate. It is further indicated that, in the case of bothstratified charge lean combustion and uniform lean combustion, thetemperature of the burnt gas rises as the air/fuel ratio becomessmaller. In order to achieve a higher temperature in the followingcylinders 2B, 2C, it is therefore appropriate to set a small air/fuelratio, whichever the combustion,mode. The relationship between the loadregion of the engine and the combustion mode that is adopted will bedescribed later.

The valve deactivating mechanism controller 54 and intake air quantitycontroller 55 have the same function as the valve deactivating mechanismcontroller 42 and intake air quantity controller 43 in FIG. 3.

The combustion controller 56 comprises fuel injection controller 57 andignition controller 58.

In this combustion controller 56, control of the combustion condition(fuel control and ignition control) is altered in accordance with themode that is set by the mode setting means 53; it also performsappropriate changeover of the combustion mode in the preceding cylinders2A, 2D and following cylinders 2B, 2C.

Specifically, when the stratified charge lean combustion mode isselected by the mode setting means 53, the fuel injection rate iscontrolled such that the air/fuel ratio in respect of the precedingcylinders 2A, 2D is a lean air/fuel ratio greater than thestoichiometric air/fuel ratio and is preferably a lean air/fuel ratiogreater than substantially twice the stoichiometric air/fuel ratio(A/F≈30) and an injection time is set such that layering of the mixtureis produced by injection of the fuel in the compression stroke and anignition time is set such that forced ignition is performed in thevicinity of the compression top dead center.

On the other hand, in regard to the following cylinders 2B, 2C, fuel issupplied in respect of the burnt gas of lean air/fuel ratio introducedfrom the preceding cylinders 2A, 2D and the fuel injection rate iscontrolled such that substantially the stoichiometric air/fuel ratio isproduced during combustion in the following cylinders 2B, 2C. Thus, inthis special operating mode, when the temperature in the followingcylinders 2B, 2C is comparatively high, the compression self-ignitionmode is selected and an injection time is set such that a uniformmixture is produced by the fuel injection in the intake stroke andforced ignition is deactivated so that. compression self-ignition may beperformed. Also, when the temperature in the following cylinders 2B, 2Cis comparatively low, so that the forced ignition mode is selected, theinjection time is set such that fuel is injected in the compressionstroke and the ignition time is set such that forced ignition isperformed at a prescribed time in the vicinity of the compression topdead center. Also, when the forced ignition mode is selected asdescribed above, the combustion mode in the preceding cylinders 2A, 2Dis shifted from the stratified charge lean combustion mode to theuniform lean combustion mode and the fuel injection rate is controlledsuch as to produce in respect of the preceding cylinders 2A, 2D a leanair/fuel ratio that is larger than the stoichiometric air/fuel ratio,preferably a lean air/fuel ratio of substantially twice, or less, thestoichiometric air/fuel ratio and the injection time is set such that auniform mixture is produced by uniform dispersion by injection of thefuel in the intake step and the ignition time is set such that forcedignition is performed in the vicinity of the compression top deadcenter.

In the special operating mode, if, with increase in engine load, a shiftis effected from the stratified charge lean combustion mode to theuniform lean combustion mode (from A2 to A3) in response to increase inthe total injection rate of fuel in respect of the preceding cylinders2A, 2D and following cylinders 2B, 2C, the fuel injection rate etc iscontrolled such as to produce a smaller value of the air/fuel ratio thanin the case of stratified charge lean combustion as described above(stratified charge lean combustion mode) in respect of the precedingcylinders 2A, 2D and an injection time is set such that a uniformmixture is produced by uniform dispersion by injection of the fuel inthe intake stroke and an ignition time is set such that forced ignitionis performed in the vicinity of the compression top dead center. Incontrast, in respect of the following cylinders 2B, 2C, the aforesaidcompression self-ignition mode is selected and, in the same way asdescribed above, an ignition time is set such that a uniform mixture isproduced by injection of fuel in the intake step and forced ignition isdeactivated so that compression self-ignition may be performed.

Specifically, as described above, when a shift takes place to theuniform lean combustion mode, in respect of the preceding cylinders 2A,2D, combustion is effected with an air/fuel ratio that is smaller thanthe air/fuel ratio in the aforesaid stratified charge lean combustion(stratified charge lean combustion mode) i.e. in a uniform leancondition that is enriched compared with combustion in the stratifiedlean condition. This air/fuel ratio, as described above, is a leanair/fuel ratio that is larger than the stoichiometric air/fuel ratio andis preferably a value of substantially twice the stoichiometric air/fuelratio or smaller than this; that is, in terms of the air excess ratio λ,is set to at least 1 and preferably less than 2.

Control when the ordinary operating mode is set is the same as in theother embodiments described above.

Next, the operation of a device according to this embodiment isdescribed.

In a device according to this embodiment also, in the operating region Aof low load and low rotational speed, the device is put into a specialoperating mode in which combustion is performed with two cylinders in aconnected condition and, in the operating region of high load or highrotational speed, the device is put into the ordinary operating mode inwhich combustion is conducted with the intake ports and exhaust ports ofall the cylinders in an independent condition.

In the intermediate/low load side operating regions A301, A302 of thisspecial operating mode region, the mode setting means 53 selects acombustion mode (stratified charge lean combustion mode) in which thecombustion mode in the preceding cylinders 2A, 2D is in a stratifiedlean condition, in which fuel is injected in the compression strokewhilst the fuel injection rate is controlled such that the air/fuelratio in the preceding cylinders 2A, 2D is a lean air/fuel ratio largerthan the stoichiometric air/fuel ratio, preferably an air/fuel ratiolarger than twice the stoichiometric air/fuel ratio and ignition isconducted at a prescribed ignition time such that stratified charge leancombustion is performed (see FIG. 5).

That is, in the intermediate/low load side operating regions A1, A2 ofthe engine, by conducting stratified charge lean combustion in thepreceding cylinders 2A, 2D, combustion can be conducted under super-leanconditions in these intermediate/low load side operating regions A1, A2where, comparatively speaking, torque is not required, thereby enablingfuel cost performance to be improved. Furthermore, when combustion isconducted in the stratified lean condition, compression self-ignition inthe following cylinders 2B, 2C can be achieved smoothly and in stablefashion, since the burnt gas is at a higher temperature than in the casewhere combustion is conducted in the uniform lean condition.

Also, in the period in which the intake stroke of the precedingcylinders 2A, 2D and the exhaust stroke of the following cylinders 2B,2C overlap, combustion is conducted while controlling the fuel injectionrate such as to provide the stoichiometric air/fuel ratio by supplyingfuel to this burnt gas while the burnt gas exhausted from the precedingcylinders 2A, 2D is being introduced into the following cylinders 2B, 2Cthrough the gas passages 22..

In this case, as a rule, compression self-ignition mode is selected and,as shown in FIG. 5, fuel is injected in the following cylinders 2B, 2Cin the intake stroke; the interior of the combustion chamber thenassumes a suitably high temperature, high pressure condition in thevicinity of the top dead center of the compression stroke andcompression self-ignition is thereby performed in a fully satisfactoryfashion. The operation and effects which are thereby obtained are asalready described in the description of the other embodiments.

However, as described above, the water temperature sensor 49 detects theengine temperature, in particular the engine temperature of thefollowing cylinders 2B, 2C constantly or at least in the low loadoperating region A301 of the engine and if the result of the detectionby this water temperature sensor 49 is lower than the prescribedtemperature at which stable compression self-ignition can be achieved inthe following cylinders 2B, 2C, the temperature status identifier 52concludes that compression self-ignition in the following cylinders 2B,2C will be difficult and the mode setting means 53 therefore effects ashift from the compression self-ignition mode to the forced ignitionmode and, as shown in FIG. 23, fuel is injected into the followingcylinders 2B, 2C in the compression stroke and combustion is performedby performing forced ignition at a prescribed ignition time.

At this point, the mode setting means 53 changes over the combustionmode of the preceding cylinders 2A, 2D from the stratified charge leancombustion mode to the uniform combustion mode. That is, even in the lowload operating region A301 of the special operating mode region A, ifthe temperature status identifier 52 concludes that compressionself-ignition will be difficult in the following cylinders 2B, 2C, themode setting means 53 shifts the combustion mode in the precedingcylinders 2A, 2D from stratified charge lean combustion to uniform leancombustion and the air/fuel ratio in the preceding cylinders 2A, 2D isset to an air fuel ratio value that is smaller than that duringstratified charge lean combustion; that is, the fuel injection rate iscontrolled and fuel is injected in the intake stroke such that a leanair/fuel ratio that is larger than the stoichiometric air/fuel ratio,preferably an air/fuel ratio of substantially twice the stoichiometricair/fuel ratio or less than this, is produced in the preceding cylinders2A, 2D, while effecting enrichment compared with the stratified leancondition. When fuel is thus injected in the intake stroke, it isuniformly dispersed in the combustion chamber by the gaseous current,resulting in a uniform fuel distribution. Ignition is thereby conductedat the prescribed ignition time and combustion is performed in a uniformlean condition (see FIG. 23).

That is, in the low load side operating region A301 of the engine, ifthe engine temperature of the following cylinders 2B, 2C is lower thanthe prescribed temperature, compression self-ignition cannot beperformed in the following cylinders 2B, 2C in stable fashion, socombustion is effected by forced ignition in the following cylinders 2B,2C and introduction of high temperature burnt gas into the followingcylinders 2B, 2C is effected with an enriched air/fuel ratio in thepreceding cylinders 2A, 2D in order that compression self-ignition canbe achieved at an early stage in the following cylinders 2B, 2C.

In this way, by enriching the air/fuel ratio of the preceding cylinders,the burnt gas thereof can be raised in temperature, so the temperaturein the following cylinders is raised and compression self-ignition inthe following cylinders can be achieved at an early stage. While itmight be thought that fuel costs would be adversely affected by loweringthe air/fuel ratio in the preceding cylinders, in fact, deterioration offuel costs is suppressed by changing over the combustion mode fromstratified charge lean combustion to uniform lean combustion, which isof better fuel cost performance.

However, FIG. 22 shows that, even when the air/fuel ratio is low,stratified charge lean combustion makes it possible to introduce burntgas at a higher temperature into the following cylinders 2B, 2C thandoes uniform lean combustion, making it possible to achieve compressionself-ignition of the following cylinders 2B, 2C at an earlier stage inthe case of combustion in the stratified lean condition. However, inthis case, there may be concern regarding deterioration of fuel costperformance due to increase of the HC exhaust rate, so, in aiming for abalance of improvement in fuel costs performance and early achievementof compression self-ignition in the following cylinders 2B, 2C, it ispreferable to perform combustion in the uniform lean condition in thepreceding cylinders 2A, 2D, as in this embodiment.

Also, in the low load operating region A301 of the engine, as shown inFIG. 25, control is effected such that the air/fuel ratio is smallduring combustion in the preceding cylinders 2A, 2D, in response tolowering of the engine load. That is, in view of the fact that thetemperature in the following cylinders 2B, 2C becomes low as the engineload becomes lower, control is performed in such cases so as to effectfurther enrichment, by increasing the fuel injection rate. This raisesthe burnt gas temperature in the preceding cylinders 2A, 2D, so thatcompression self-ignition can be performed in a smooth and stablefashion in the following cylinders 2B, 2C without inviting deteriorationof fuel costs.

Then, with gradual increase in the engine load, in the intermediate loadoperating region A302 of the engine, combustion is effected in thepreceding cylinders 2A, 2D with a fixed air/fuel ratio in the super-leancondition and, with further increase in engine load, in the high loadside operating region A303 of the engine, the air/fuel ratio in thepreceding cylinders 2A, 2D is gradually reduced and combustion isperformed in the uniform lean condition.

Specifically, in the high load side operating region A303 of the specialoperating mode region A, the mode setting means 53 selects uniform leancombustion (uniform lean combustion mode) as the combustion mode in thepreceding cylinders 2A, 2D; the fuel injection rate is controlled andfuel injection is performed in the intake stroke such that, albeit theair/fuel ratio in the preceding cylinders 2A, 2D is a value smaller thanthat in the stratified charge lean combustion mode i.e. is richer thanin the stratified lean condition, the air/fuel ratio in the precedingcylinders 2A, 2D is a lean air/fuel ratio larger than the stoichiometricair/fuel ratio and preferably substantially twice the stoichiometricair/fuel ratio or an air/fuel ratio smaller than this. Thus, when fuelis injected in the intake step, the distribution of the fuel becomesuniform by uniform dispersal within the combustion chamber by thegaseous current. Thus, ignition is performed at the prescribed ignitiontime and combustion is conducted under uniform lean conditions (see FIG.24).

Specifically, as engine load becomes higher, in general, more torque isrequired and the fuel injection rate is increased. As this fuelinjection rate is increased, the air/fuel ratio naturally decreases i.e.is enriched to a range where ignition under uniform lean conditionsbecomes possible. When the air/fuel ratio is decreased to a range atwhich ignition becomes possible under such uniform lean conditions, thetemperature of the burnt gas becomes lower than in the case ofstratified charge lean combustion so a shift is brought about to uniformlean combustion, which offers excellent fuel cost performance.

Thus, in the high load side operating region A303 in which high torqueis required, in general the engine temperature becomes high, increasingthe risk of knocking, so, in such cases, a changeover is effected tocombustion under uniform lean conditions in which the burnt gastemperature is lower than in the case of combustion under stratifiedlean conditions, so occurrence of knocking can thereby be effectivelysuppressed. Also, in combustion under uniform lean conditions, fuel costperformance for the same load and same air/fuel ratio is better than inthe case of combustion under stratified lean conditions, so aconsiderable improvement in the fuel cost performance can be achieved byadopting combustion under uniform lean conditions in the high load sideoperating region A303 where fuel injection is increased due to hightorque being demanded.

Also, in the high load side operating region A303, the air/fuel ratioduring combustion becomes smaller in the preceding cylinders 2A, 2D asthe engine load becomes higher, so knocking can be more effectivelysuppressed since the EGR is also increased with this increase in fuelinjection rate.

FIG. 26 shows the drive/control system according to yet a furtherembodiment of the present invention. In this embodiment, the ECU 60comprises operating condition identifier 61, temperature statusidentifier 62, mode setting means 63, valve stop mechanism controller64, intake air quantity controller 65, fuel controller 66 and ignitioncontroller 67.

The operating condition identifier 61 comprises a map for controlpurposes as shown in FIG. 4, like that of the operating conditionidentifier 41 in FIG. 3, whereby it determines whether the operatingcondition is in region A or B. The temperature status identifier 62ascertains whether the water temperature (engine temperature) is a lowtemperature, below a prescribed value, or a high temperature, above theprescribed temperature.

In accordance with the determination by the operating conditionidentifier 61, the mode setting means 63 selects the special operatingmode in the region A and selects the ordinary operating mode in theregion B.

The valve stop mechanism controller 64 and intake air quantitycontroller 65 have the same functions as the valve stop mechanismcontroller 42 and intake air quantity controller 43 in FIG. 2.

Also, the fuel controller 66 and ignition controller 67 perform controlof the combustion condition (control of fuel and control of the ignitiontime) in accordance with the mode set by the mode setting means 63.

That is, when the special operating mode is set, the fuel injection ratein respect of the preceding cylinders 2A, 2D is controlled such as toproduce a lean air/fuel ratio greater than the stoichiometric air/fuelratio, preferably substantially twice the stoichiometric air/fuel ratioor more than this and the injection time is set such as to achievestratified charge combustion by fuel injection in the compression strokeand, furthermore, an ignition time is set whereby forced ignition isperformed in the vicinity of the compression top dead center. On theother hand, the fuel injection rate, fuel injection time and the fuelinjection valve 9 that performs this fuel injection are set such as tocontrol the fuel ignition rate so as to produce the stoichiometricair/fuel ratio by supplying fuel to the burnt gas of lean air/fuel ratioand to supply this fuel by the preceding cylinders 2A, 2D in respect ofthe following cylinders 2B, 2C, i.e. so as to inject fuel of an amountappropriate to the following cylinders directly into the combustionchambers 4 during the exhaust stroke of the preceding cylinders 2A, 2Dby means of the fuel injection valve 9 of the preceding cylinders 2A,2D; in addition, forced ignition is deactivated so as to performcompression ignition in the following cylinders 2B, 2C.

In this way, in this embodiment, the fuel injection valves 9 of thepreceding cylinders 2A, 2D, to be described, of the fuel injectionvalves 9 of the cylinders 2A to 2D are arranged to function as thesecond fuel injection means according to the present invention and thefuel controller according to the present invention is constituted by thefuel controller 46 and ignition controller 47 etc.

Control when the ordinary operating mode is set is the same as in theother embodiments, already described.

Next, the operation of a device according to this embodiment isdescribed with reference to FIG. 27.

In the operating region A of low load, low rotational speed, the specialoperating mode is set and the condition in which two cylinders areconnected (see FIG. 6) is produced.

In this condition, new air is introduced from the intake passage 15 inthe respective intake strokes to the preceding cylinders 2A, 2D, fuel isinjected in the compression stroke while controlling the fuel injectionrate such that the air/fuel ratio in the preceding cylinders 2A, 2D is alean air/fuel ratio greater than the stoichiometric air/fuel ratio andstratified charge combustion is performed with the lean air/fuel ratioby performing ignition at the prescribed ignition time.

After this, fuel for the following cylinders is directly injected intothe combustion chambers 4 of the preceding cylinders 2A, 2D during theperiod in which the exhaust stroke of the preceding cylinders 2A, 2D andthe intake stroke of the following cylinders 2B, 2C overlap and the fuelinjection rate is controlled such that the burnt gas of the leanair/fuel ratio produces the stoichiometric air/fuel ratio (F₂ in FIG.27); this burnt gas, containing fuel, is introduced into the followingcylinders 2B, 2C (white arrow in FIG. 27) through the inter-cylinder gaspassages 22 whilst being exhausted from the preceding cylinders 2A, 2D.Compression self-ignition is then performed in the following cylinders2B, 2C by rise of the pressure and temperature within the combustionchambers in the vicinity of the top dead center of the compressionstroke. Since the high temperature burnt gas that is exhausted from thepreceding cylinders 2A, 2D is then immediately introduced into thefollowing cylinders 2B, 2C through the short inter-cylinder gas passages22, the temperature within the combustion chambers in the followingcylinders 2B, 2C becomes high in the intake stroke and, by further risein pressure and temperature from this condition in the compressionstroke, the temperature within the combustion chambers is raised to sucha degree as to enable thoroughly satisfactory self-ignition of themixture in the vicinity of the top dead center in the latter period ofthe compression stroke.

The burnt gas after combustion in the following cylinders 2B, 2C is thenexhausted to the exhaust passage 20, which is provided with a three-waycatalyst 24.

In this way, due to the super-lean combustion in the preceding cylinders2A, 2D and the combustion by compression self-ignition in the followingcylinders 2B, 2C, the operation and benefits already described areobtained in the same way as in the case of the other embodiments.

In particular, with this embodiment, since injection of the fuel in anamount appropriate to the following cylinders is effected in thepreceding cylinders 2A, 2D, the mixed gas and high temperature burnt gasare thoroughly mixed between the burnt gas being exhausted from thepreceding cylinders 2A, 2D and being introduced into the followingcylinders 2B, 2C and so are uniformly distributed; a uniform mixturedistribution condition and mixture temperature satisfying the idealcompression self-ignition condition are therefore obtained.Self-ignition performance in the following cylinders 2B, 2C is thereforeare improved and excellent combustion achieved.

Although, in the above embodiment, it was arranged for fuel in an amountappropriate to the following cylinders to be injected into the precedingcylinders 2A, 2D in the special operating mode by using the fuelinjection valves 9 of the cylinders, it would be possible for example asshown in FIG. 28 to arrange for a special-purpose fuel injection valve 9a (second fuel injection means) to be provided at some point along theinter-cylinder gas passages 22 and for fuel to be supplied to the burntgas prior to its introduction into the following cylinders 2B, 2C afterthe exhaustion from the preceding cylinders 2A, 2D in an amountappropriate to the following cylinders. The essential point is that itshould be possible to obtain a mixing effect between the mixture andhigh temperature burnt gas by supplying fuel in an amount appropriate tothe following cylinders to the burnt gas of the preceding cylinders 2A,2D prior to introduction thereof into the following cylinders 2B, 2C andthat, as a result, the self-ignition performance is improved; theinjection timing of the fuel in an amount appropriate to the followingcylinders therefore does not matter so long as the arrangement is suchthat this fuel in an amount appropriate to the following cylinders canbe supplied to the burnt gas with such a timing.

However, in the case of a direct injection engine, as in the embodimentdescribed above, it is possible to inject the fuel during the exhauststroke of these cylinders by using the fuel injection valves 9 of thepreceding cylinders 2A, 2D, so, in this case, the provision of aspecial-purpose fuel injection valve for supplying fuel for thefollowing cylinders as in FIG. 28 is unnecessary, so there is theadvantage that a straightforward construction can be achieved in whichthe basic construction of a direct injection engine is utilized withoutmodification.

In a construction in which, as in FIG. 28, a special-purpose fuelinjection valve 9 a is provided for supplying fuel in an amountappropriate to the following cylinders, for example instead of providinga fuel injection valve 9 in each cylinder, it would be possible toprovide a fuel injection valve 9 b in a branched intake passage 16 asshown by the broken line in this Figure and to perform fuel injection tothe cylinders 2A to 2D in the case of the ordinary operating mode orfuel injection to the preceding cylinders 2A, 2D in the case of thespecial operating mode by port injection.

Also, when fuel in an amount appropriate to the following cylinders issupplied in respect of the burnt gas in the preceding cylinders, aconstruction may be adopted. wherein a fuel injection valve capable ofin-cylinder injection is provided at least in respect of the precedingcylinders, so fuel injection may be achieved by port injection forexample by providing a fuel injection valve in the intake passages inrespect of the following cylinders.

Also, although, in the foregoing embodiments, it was arranged to supplyfuel in an amount appropriate to the following cylinders into thepreceding cylinders 2A, 2D in all cases in this special operating mode,it would be possible to ascertain the degree of self-ignition capabilityin the following cylinders 2B, 2C and, in the case of an operatingcondition in which this self-ignition capability is high, to performcompression ignition by supplying fuel in the intake stroke of thefollowing cylinders 2B, 2C after introduction of the burnt gas of thepreceding cylinders 2A, 2D into the following cylinders 2B, 2C as shownin FIG. 29 (this may be called the first injection mode), or, in thecase of an operating condition in which the self-ignition capability islow, to supply fuel in amount appropriate to the following cylindersinto the preceding cylinders 2A, 2D as described in the above embodiment(this may be called the second injection mode). For example, in anoperating condition in which the temperature of the engine as determinedby the temperature status identifier 42 is below a specifiedtemperature, preferably it is arranged for combustion to besatisfactorily conducted by raising the self-ignition capability in thefollowing cylinders 2B, 2C by performing the second injection mode.

It should be noted that the determination of the degree of self-ignitioncapability may be performed for example in the fuel controller 66 inaccordance with information relating to the operating condition andapart from being determined in accordance with engine temperature asdescribed above, could be determined in accordance with enginerotational speed or engine load etc. For example, it may be assumedthat, in particular in a very low load region of the operating region A,the fuel injection rate will become low and the capability forself-ignition will diminish. Combustion can therefore be achieved bycompression self-ignition in the following cylinders 2B, 2C in a fullysatisfactory manner in such circumstances also, by arranging to performthe second ignition mode.

Also, although, in the above embodiments, the region A of low speed andlow load is designated as a special operating mode region and combustionby compression self-ignition is arranged to be performed in all cases inthe following cylinders in this special operating mode, it could bearranged to perform combustion by forced ignition in part of this regionA.

FIG. 30 and FIG. 31 show yet a further embodiment of the presentinvention.

In the overall construction shown in FIG. 30, the main engine body 1,the intake/exhaust ports and the intake/exhaust passages connectedtherewith, the inter-cylinder gas passages and the intake/exhaust valvesetc that open and close the intake/exhaust ports are practically thesame as those shown in FIG. 1 or FIG. 8. Furthermore, in thisembodiment, it is arranged for the opening/closing times of the valvesto be varied in accordance with conditions by means of cam phase varyingmechanisms 33 a, 34 a provided in a valve moving mechanism for theintake/exhaust valves and for these to be closed off by a valvedeactivating mechanism 35.

The cam phase varying mechanisms 33 a, 34 a are previously knownmechanisms that vary the rotational phase of camshafts 33, 34 withrespect to the rotational phase of the crankshaft. As shown in FIG. 1,the cam phase varying mechanism 33 a is provided on the camshaft 33 andthe cam phase varying mechanism 34 a is provided on the camshaft 34,these being independently controlled (see FIG. 31). Consequently, theopening/closing time of the preceding cylinder intake valves 31 andfollowing cylinder intake valves (first intake valves) 31 a that areopened and closed by rotation of the camshaft 33 is varied in advance orin retardation overall by the cam phase varying mechanism 33 a. In thesame way, the opening/closing time of the burnt gas introduction valve(second intake valve) 31 b, the following cylinder exhaust valve 32, thepreceding cylinder exhaust valve (first exhaust valve) 32 a and burntgas exhaust valve (second exhaust valve) and 32 b that are opened andclosed by rotation of the cam shaft 34 is varied in advance or inretardation overall by the cam phase varying mechanism 34 a.

The ECU 70 in FIG. 31 comprises operating condition identifier 71, valvestop mechanism controller 720, intake air quantity controller 73, fuelcontroller 74 and cam phase controller 77.

The operating condition determining means 71 ascertains whether theoperating condition is in the operating region A on the low load, lowrotational speed side shown in FIG. 32 (in which the engine load is lessthan T1 and the engine rotational speed is less than r1) or is in theoperating region B on the high load side or high rotational speed side(in which the engine load exceeds T1 or the engine rotational speedexceeds r1) and furthermore, if the operating condition is in theoperating region A, ascertains whether it is in the comparatively lowload, low rotational speed region A401 thereof or is in thecomparatively high load, high rotational speed region A402 thereof.Then, under the prescribed condition (for example condition in which theengine is fully warmed-up), operation is conducted in the specialoperating mode in which the cylinders are put in a two-cylinderconnected condition in the operating region A and operation is conductedin the ordinary operating mode in which the cylinders are in anindependent condition in the operating region B.

The valve deactivating mechanism controller 72 and intake air quantitycontroller 73 have the same function as the valve deactivating mechanismcontroller 42 and intake air quantity controller 43 in FIG. 3. Also, thecombustion controller 74 comprises fuel injection controller 75 andignition controller 76 and performs fuel injection and ignition controlin accordance with the operating regions A, B in substantially the sameway as the combustion controller 44 in FIG. 3.

The cam phase controller 77 controls the cam phase varying mechanisms 33a, 34 a in accordance with the results of the determination by theoperating condition identifier 71. The details of the control will bedescribed later, but for example in the special operating mode the camphase varying mechanism 34 a is controlled in a direction such as toadvance the phase of the cam 27 in the comparatively low load, lowrotational speed region (region A401 in FIG. 32), so that it is set suchthat the opening/closing times of the burnt gas exhaust valve 32 b,burnt gas introduction valve 31 b and following cylinders' exhaustvalves 32 that are actuated by rotation of the camshaft 34 all takeplace at an early stage. In contrast, in the comparatively high load,high rotational speed region (region A402 in FIG. 32), the cam phasevarying mechanism 34 a is controlled in a direction such as to delay thephase of the cam 27, so that it is set such that the opening/closingtimes of the burnt gas exhaust valve 32 b, burnt gas introduction valve31 b and following cylinders' exhaust valves 32 that are actuated byrotation of the camshaft 34 are all delayed. It should be noted that thecam phase varying mechanisms of 33 a, 34 a act on the valves whilst theyare activated, so that a valve that has been put into the deactivatedcondition by the valve deactivating mechanism 35 remains in thedeactivated condition irrespective of control operation of the cam phasevarying mechanisms 33 a, 34 a.

Next, the operation of this embodiment will be described with referenceto FIG. 33 and FIG. 34.

In the device of this embodiment also, in the low load, low rotationalspeed operating region A, the device is put into the special operatingmode in which combustion is conducted in the two-cylinder connectedcondition and, in the high load or high rotational speed operatingregion B, the device is put into the ordinary operating mode in whichcombustion is conducted with the intake ports and exhaust ports of therespective cylinders in an independent condition. When in the specialoperating mode, combustion is conducted in the preceding cylinders witha super-lean air/fuel ratio and in the following cylinders combustion isconducted by compression self-ignition.

However, even in the special operating mode, if the cylinder temperatureof the following cylinders 2B, 2C is low, so that they are in acondition in which compression ignition is difficult, a changeover ofthe ignition of the following cylinders 2B, 2C to forced ignition by aspark plug 7 is effected. Also in the contrary situation in which thecylinder temperature of the following cylinders 2B, 2C is too high,resulting in a condition in which abnormal combustion such as knockingoccurs, changeover is effected from the special operating mode to theordinary operating mode. In both cases, the benefit of improvement infuel costs etc is suppressed compared with the special operating modeusing compression ignition. It is therefore desirable to expand theoperating region that is suited to compression ignition in the specialoperating mode, in order to obtain such benefits to a greater extent.

The opening/closing times of the intake valve are set as follows inorder to expand the operating region that is suited to compressionignition in the special operating mode.

FIG. 33 shows in detail the opening/closing times etc of theintake/exhaust valves and is a diagram showing the opening/closing timeof the preceding cylinders' intake valves 31 and the burnt gas exhaustvalves 32 b of the preceding cylinders 2A, 2D in the special operatingmode and the opening/closing times of the burnt gas introduction valves31 b of the following cylinders 2B, 2C and the following cylinders'exhaust valves 32. FIG. 33(a) is the case of the comparatively low load,low rotational speed region (region A401 in FIG. 32) of the operatingcondition in which the special operating mode is conducted and FIG.33(b) is likewise the case of the comparatively high load, highrotational speed region (region A402 in FIG. 32). In these Figures, thehorizontal axis shows the crank angle and T is the top dead center (TDC)while B is the bottom dead center (BDC). The interval between T and B is108° CA. Also, the upper section shows the preceding cylinders 2A, 2Dwhile the lower section shows the following cylinders 2B, 2Ccorresponding thereto. Also, the band-shaped portions indicate the openperiods of the valves. The white arrow from the upper section to thelower section indicates a condition in which the burnt gas from thepreceding cylinders 2A, 2D is fed into the following cylinders 2B, 2C,with the exhaust stroke of the preceding cylinders 2A, 2D and the intakestroke of the following cylinders 2B, 2C overlapping.

FIG. 33(a) shows in the upper section thereof the open period 80 of theburnt gas exhaust valve in which the burnt gas exhaust valves 32 b ofthe preceding cylinders 2A, 2D are open and the open period 81 (shownshaded) of the preceding cylinders' intake valves in which the precedingcylinders' intake valves 31 are open. In the lower section thereof, itshows the open period 82 of the following cylinders' exhaust valves inwhich the following cylinders' exhaust valves 32 of the followingcylinders 2B, 2C are open and the open period 83 of the burnt gasintroduction valves in which the burnt gas introduction valves 31 b areopen. The open period 81 of the preceding cylinders' intake valves isset from about 10° CA before the TDC to about 55° CA after the BDC(total about 245° CA). This is a typical set value in prior art engines.In contrast, the open period 83 of the burnt gas introduction valves isset from about 45° CA before the TDC to substantially the BDC (totalabout 225° CA). Also, the open period 82 of the following cylinders'exhaust valves is set to about 80° CA before the BDC to about 25° CAbefore the TDC (total about 235° CA).

That is, the interval between the bottom dead center 96 of the followingcylinders' intake stroke and the open period 97 of the burnt gasintroduction valves (about 0° CA) is set to be shorter than the interval(about 55° CA) between the bottom dead center 92 of the precedingcylinders' intake stroke and the closed period 93 of the precedingcylinders' intake valves. Also the open period 83 of the burnt gasintroduction valves is shorter than the open period 81 of the precedingcylinders' intake valves and the open period 82 of the followingcylinders' exhaust valves. Thus, since, in the following cylinders 2B,2C, the open period 83 of the burnt gas introduction valves is set to beshort and the closed period 97 of the burnt gas introduction valves isset to be earlier and at a time close to the BDC, the period for whichthe burnt gas introduction valve 31 b is open during ascent of thepiston 3 beyond the BDC is non-existent or short. The effectivecompression ratio is therefore increased, approaching the geometricalcompression ratio. With this increase in the effective compressionratio, the cylinder temperature of the following cylinders 2B, 2C tendsto rise, causing the compression ignition capability to increase. Itshould be noted that what is introduced into the following cylinders 2B,2C at this point is not new air throttled by the multi-linked throttlevalves 17 but burnt gas exhausted from the preceding cylinders 2A, 2D,which easily flows into the combustion chambers 4, so these aresatisfactorily filled even if the burnt gas introduction valves 31 b areclosed at an early stage.

Also, the open period 80 of the burnt gas introduction valves is setfrom about 80° CA before the BDC to about 25° CA before the TDC (totalabout 235° CA). The closed period 19 of the burnt gas exhaust valves istherefore set advanced by about 25° CA compared with the closed period97 of the burnt gas introduction valves (corresponding to the TDC in thepreceding cylinders 2A, 2D). This is earlier than the typical set valueof a prior art engine (about 50° CA after the TDC). By closing the burntgas exhaust valves 32 b at an early stage, the amount of internal EGR ofthe preceding cylinders 2A, 2D is increased, thereby increasing thetemperature of the burnt gas that is introduced into the followingcylinders 2B, 2C.

As described above, the compression ignition capability is improved byraising the cylinder temperature of the following cylinders 2B, 2C byincreasing the effective compression ratio of the following cylinders2B, 2C and raising the temperature of the burnt gas that is introducedinto the following cylinders 2B, 2C. In this way, the operating regionin which combustion by compression ignition can be performed can beexpanded further into the low load region.

The upper section of FIG. 33(b) shows the open period 100 of the burntgas exhaust valves for which the burnt gas exhaust valves 32 b of thepreceding cylinders 2A, 2D are open and the open period 101 (indicatedby shading) of the preceding cylinders' intake valves, for which thepreceding cylinders' intake valves 31 are open. The lower sectionthereof shows the open period 102 of the following cylinders' exhaustvalves for which the following cylinders' exhaust valves 32 of thefollowing cylinders 2B, 2C are open and the open period 103 of the burntgas introduction valves for which the burnt gas introduction valves 31 bare open. The open period 101 of the preceding cylinders' intake valvesis set in the same way as the open period 81 of the preceding cylinders'intake valves of FIG. 33(a). In contrast, regarding the open period 100of the burnt gas exhaust valves, the open period 102 of the followingcylinders' exhaust valves and the open period 103 of the burnt gasintroduction valves (respectively shown by white band lines), these areset such that the lengths of the respective periods are equal to theopen period 80 of the burnt gas exhaust valves, open period 82 of thefollowing cylinders' exhaust valves and open period 83 of the burnt gasintroduction valves, while only the opening/closing times of each valveare set to be delayed by about 45° CA overall. This is achieved bydelaying the phase of the camshaft 34 by about 45° CA by means of thecam phase controller 49, as described above (see FIG. 31).

The closed period 114 of the burnt gas introduction valves of FIG. 33(b)is delayed by about 45° CA from the closed period 97 of the burnt gasintroduction valves of FIG. 33(a). As a result, even during the rise ofthe pistons 3 in the following cylinders 2B, 2C after having passed thefollowing cylinders' intake stroke bottom dead center 113, the burnt gasintroduction valves 31 b are open for about 45° CA. The effectivecompression ratio of the following cylinders 2B, 2C is therefore reducedfrom that in the case of FIG. 33(a) and the cylinder temperature of thefollowing cylinders 2B, 2C falls.

Also, the closed period 112 of the burnt gas exhaust valves is delayedby about 45° CA from the closed period 90 of the burnt gas exhaustvalves of FIG. 33(a), being set to about 20° CA after the top deadcenter 111 of the preceding cylinders' exhaust stroke. Consequently, theinternal EGR of the preceding cylinders 2A, 2D is reduced from that inthe case of FIG. 33(a) and the temperature of the burnt gas that isintroduced into the following cylinders 2B, 2C falls.

As described above, due to the lowering of the effective compressionratio of the following cylinders 2B, 2C and the fall in the temperatureof the burnt gas introduced into the following cylinders 2B, 2C, thecylinder temperature of the following cylinders 2B, 2C falls, preventingoccurrence of abnormal combustion such as knocking. In this way, theoperating region in which combustion can be performed by compressionignition is extended further into the high load region.

Thus, in the special operating mode, the opening/closing times of theintake/exhaust valves are set in a direction such as to raise thecylinder temperature thereof in an operating region in which thecylinder temperature of the following cylinders 2B, 2C is comparativelylow and in a direction such as to lower the cylinder temperature in anoperating region in which this is comparatively high. The operatingregion in which combustion can be performed appropriately by compressionignition in the following cylinders 2B, 2C can thereby be expanded,making it possible to further promote the effect of improvement of fuelcosts and exhaust gas cleansing.

FIG. 34 is a diagram given in explanation of the opening/closing timesof the preceding cylinders' intake valves 31 and the burnt gas exhaustvalves 32 b of the preceding cylinders 2A, 2D in the ordinary operatingmode, and also of the opening/closing times of the burnt gasintroduction valves 31 b and following cylinders' exhaust valves 32 ofthe following cylinders 2B, 2C. The horizontal axis shows the crankangle, T being the top dead center (TDC) and B being the bottom deadcenter (BDC). The interval between T and B is 108° CA. Also, the uppersection shows the preceding cylinders 2A, 2D while the first sectionshows the following cylinders 2B, 2C. In the ordinary operating mode,the cylinders operate independently, so combustion is performed byforced ignition when new air is introduced, both in the case of thepreceding cylinders 2A, 2D and the following cylinders 2B, 2C. The uppersection of FIG. 34 shows the open period 120 of the burnt gas exhaustvalves in which the preceding cylinders' exhaust valves 32 a of thepreceding cylinders 2A, 2D are open and the open period 121 of thepreceding cylinders' intake valves for which the preceding cylinders'intake valves 31 are open. The bottom section shows the open period 122of the following cylinders' exhaust valves for which the bottomcylinders' exhaust valves 32 of the following cylinders 2B, 2C are openand the open period 123 of the following cylinders' intake valves forwhich the following cylinders' intake valves 31 a are open.

The exhaustion of the preceding cylinders 2A, 2D and intake by thefollowing cylinders 2B, 2C are performed by different valves than in thespecial operating mode, so this opening/closure is performed by means ofa different cam. The open period of the 120 of the preceding cylinders'exhaust valves and the open period 123 of the following cylinders'intake valves can therefore be set independently from the opening period80 of the burnt gas exhaust valves and the open period 83 of the burntgas introduction valves of FIG. 33(a). In FIG. 34, the open period 121of the preceding cylinders' intake valves and the open period 123 of thefollowing cylinders' intake valves (shown shaded) are set from about 10°CA before TDC to about 55° CA after BDC (total about 245° CA). Also, theopen period 120 of the preceding cylinders' exhaust valves and the openperiod 122 of the following cylinders' exhaust valves are set from about30° CA before BDC to about 25° CA after TDC (total about 235° CA). Theseopen periods correspond to typical settings of a prior art engine.

The open period 121 of the preceding cylinders' intake valves and theopen period 123 of the following cylinders' intake valves can be variedforwards and backwards by the cam phase varying mechanism 33 a and theopen period 120 of the preceding cylinders' exhaust valves and the openperiod 122 of the following cylinders' exhaust valves can be variedforwards and backwards by means of the cam phase varying mechanism 34 a.Consequently, the period for which the preceding cylinders' exhaustvalves 32 a and the preceding cylinders' intake valves 31 are both open(period in which the open period 120 of the preceding cylinders' exhaustvalves and the open period 121 of the preceding cylinders' intake valvesoverlap: “valve overlap”) can be varied by controlling the cam phasevarying mechanisms 33 a, 34 a. The cam phase controller 49 controls thecam phase varying mechanisms 33 a, 34 a so that the valve overlapincreases as the load increases, so that optimum combustion efficiencyis obtained in accordance with the load. The same control is performedin respect of the following cylinders 2B, 2C.

In this way, in the ordinary operating mode, output performance can beensured by effecting control to optimum intake/exhaust times inaccordance with the load and by controlling the intake air rate and fuelinjection rate such as to produce a stoichiometric air/fuel ratio or aratio richer than this.

It should be noted that the patterns shown in FIGS. 33(a) and (b) andFIG. 34 are not restrictive of the respective valve opening periods ofvalve opening times and these may be suitably varied within the scope ofthe claims. For example, the closing time 97 of the burnt gasintroduction valve of FIG. 33(a) may be set to be somewhat later thanthe bottom dead center 96 of the intake stroke of the followingcylinders. Also, the opening periods 81, 101, 121 of the intake valvesof the preceding cylinders of FIGS. 33(a) and (b) and FIG. 34 may be setto be different depending on the load (in this embodiment, this iseffected by means of the cam phase varying mechanism 33 a), or may beset to be the same as in this embodiment (in this case, the cam phasevarying mechanism 33 a need not be provided). Other specific set valuesmay also be set to suitable values in accordance with the demandedengine performance.

Also, the operating region A in the special operating mode may bedivided into three or more regions instead of into two regions A401,A402 and suitable valve opening/closing times set in accordance with therespective regions. Furthermore, it could be arranged to producecontinuous variation rather than setting in stepwise fashion usingdivided regions.

Next, a modified example of this embodiment is described with referenceto FIG. 35 to FIG. 38. In these Figures, parts which are the same as inthe case of the first embodiment are given the same reference symbolsand repeated description thereof is dispensed with.

FIG. 35 is a partial perspective view of a cam changeover mechanism 150according to this embodiment, given in respect of the burnt gasintroduction valves 31 b etc indicated by the double-dotted chain lines.A camshaft 151 is arranged above the burnt gas introduction valves 31 b.The camshaft 151 is arranged to rotate integrally with cams of threetypes, namely, a first cam 152, a second cam 154 and a third cam 156,having independent lift characteristics. A rocker arm set 160 supportedon a rocker shaft 170 is provided between these cams and a burnt gasintroduction valve 31 b. The rocker arm set 160 is an assembly of rockerarms of three types, namely, a first rocker arm 162, second rocker arm164 and third rocker arm 166. At the tip of the first rocker arm 162,there are provided a valve abutment section 163 and an adjustment screw161 for fine adjustment of the position thereof in the axial direction;the valve abutment section 163 abuts the upper end of the valve shaft offor example the burnt gas introduction valve 31 b at a suitableposition. On the second rocker arm 164 and third rocker arm 166, thereare provided springs, outside the drawing, that press these rocker armsagainst the first cam 154 and third cam 156. Consequently, when therocker arms of the rocker arm set 160 are moveable independently, asshown, the upper surfaces of the rocker arms abut the peripheries of thefirst cam 152, second cam 154 and third cam 156 and are therebyvertically rocked about the rocker shaft 170 in accordance with theshape of the cam abutment sections (radius of rotation of the respectivecams).

In the interior of the rocker arms set 160, there are provided fiveplungers, to be described, in two rows (see FIG. 36. A fourth plungerhole 204, which is one of the plunger holes thereof, is visible in FIG.35). By the movement of these plungers, the first rocker arm 162 can bemade integral with the second rocker arm 164 or third rocker arm 166i.e. can be linked for joint movement therewith. A first passage 172 forthe supply of hydraulic fluid and a second passage 174 for the supply ofhydraulic fluid are provided that conduct oil for hydraulic operation ofthe plungers within the rocker shaft 170.

The first cam 152 is a cam for valve deactivation and has a circularperipheral shape concentric with the camshaft 151. Consequently, whenthe upper surface of the first rocker arm 162 is constantly in abutmentwith the peripheral surface of the first cam 152 (the second rocker arm164 and third rocker arm 166 being detached), no rocking takes placeeven though the camshaft 151 is rotated. In other words, the burnt gasintroduction valves 31 b etc are deactivated in the closed condition.

The second cam 154 is a cam for low load (or low speed) and comprises aportion having the same peripheral shape as the first cam 152 and aportion having a peripheral shape that projects therefrom. Consequently,when the upper surface of the second rocker arm 164 is constantly inabutment with the peripheral surface of the second cam 154 (the thirdrocker arm 166, to be described, being detached), downward rocking iseffected by a prescribed amount at a prescribed crank angle as the.camshaft 151 rotates. Thus, if the first rocker arm 162 and secondrocker arm 164 are then linked for joint movement, the operation of thefirst rocker arm 162 is the same as the rocking of the second rocker arm164 by the second cam 154. In other words, the burnt gas introductionvalve 31 b is opened by a prescribed amount at a prescribed time.

The third cam 156 is a cam for high load (or high speed) and comprises aportion having the same peripheral shape as the second cam 154 and aportion having a peripheral shape that projects therefrom. Consequently,when the upper surface of the third rocker arm 166 is constantly incontact with the peripheral surface of the third cam 156, rocking takesplace downwards by a prescribed amount at a prescribed crank angle asthe camshaft 151 rotates. Thus, if the first rocker arm 162 and thirdrocker arm 166 are then linked for joint movement, the operation of thefirst rocker arm 162 is the same as the rocking of the third rocker arm166 by the third cam 156. In other words, the burnt gas introductionvalve 31 b is opened by a prescribed amount at a prescribed time (thevalve opening period includes the valve opening period when only thesecond rocker arm 164 is linked with the first rocker arm 162).

FIG. 36 is a diagram showing the operation of the five plungers providedin the interior of the rocker arm set 160. FIG. 36(a) shows thecondition in which the first rocker arm 162 is detached from the secondrocker arm 164 and third rocker arm 166; FIG. 36(b) shows the conditionin which the first rocker arm 162 is linked with the second rocker arm164 only; and FIG. 36(c) shows the condition in which the first rockerarm 162 is linked with the second rocker arm 164 and third rocker arm166.

A first plunger hole 201 and a second plunger hole 204 are provided inthe interior of the first rocker arm 162. The first plunger hole 201 isa concave recess of circular cross section opening towards the secondrocker arm 164. At the bottom of the first plunger hole 201, a firsthydraulic fluid inlet 173 is provided, leading from a first passage 172for supply/discharge of hydraulic fluid. A first plunger 181 ofcylindrical shape is fitted into the first plunger hole 201. The firstplunger 181 slides smoothly within the first plunger hole 201 whilesealing the hydraulic fluid that is fed by the first hydraulic fluidinlet 173 at its peripheral surface. The overall length of the firstplunger 181 is shorter than the depth of the first plunger hole 201.

The fourth plunger hole 204 is a through-hole that communicates with thesecond rocker arm 164 and the third rocker arm 166. A cylindrical fourthplunger 184 is fitted into the fourth plunger hole 204. The overalllength of the fourth plunger 184 is the same as the depth of the fourthplunger hole 204 (plate thickness of the first rocker arm 162). Thefourth plunger 184 slides smoothly within the fourth plunger hole 204.

A second plunger hole 202 and a fifth plunger hole 205 are provided inthe interior of the second rocker arm 164. The second plunger hole 202is a concave recess of circular cross section opening towards the firstrocker arm 162 and is of a diameter equal to that of the first plungerhole 201. An air escape hole 206 is provided at the bottom of the secondplunger hole 202 to maintain the gaseous pressure of the interior atatmospheric pressure while allowing oil leakage to escape. A secondplunger 182 of cylindrical shape having a bottom and of externaldiameter equal to that of the first plunger 181 is fitted into thesecond plunger hole 202. The second plunger 182 slides smoothly withinthe second plunger hole 202. The overall length of the second plunger182 is equal to the depth of the second plunger hole 202. The end of thesecond plunger 182 that abuts the first plunger 181 is formed inspherical shape. A second plunger spring 187 is provided in a recess onthe inside of the second plunger 182, so that the second plunger 182 isconstantly biased towards the first plunger 181.

The fifth plunger hole 205 is a concave recess of circular cross sectionopening towards the first rocker arm 162 and is of diameter equal tothat of the fourth plunger hole 204. An air escape hole 207 is providedat the bottom of the fifth plunger hole 205 to maintain the gaseouspressure of the interior at atmospheric pressure while allowing oilleakage to escape. A fifth plunger 185 of cylindrical shape having abottom and of external diameter equal to that of the fourth plunger 184is fitted into the fifth plunger hole 205. The fifth plunger 185 slidessmoothly within the fifth plunger hole 205. The overall length of thefifth plunger 185 is shorter than the depth of the fifth plunger hole205. The end of the fifth plunger 185 that abuts the fourth plunger 184is formed in spherical shape. A fifth plunger spring 189 is provided ina recess on the inside of the fifth plunger 185, so that the fifthplunger 185 is constantly biased towards the fourth plunger 184.

A third plunger hole 203 is provided within the third rocker arm 166.The third plunger hole 203 is a concave recess of circular cross sectionopening towards the first rocker arm 162 and is at the same diameter asthe first plunger hole 204. At the bottom of the third plunger hole 203,a second hydraulic fluid inlet 175 is provided, leading from a secondpassage 174 for supply/discharge of hydraulic fluid. A third plunger 183of cylindrical shape equal in diameter to the fourth plunger 184 isfitted into the third plunger hole 203. The third plunger 183 slidessmoothly within the third plunger hole 203 while sealing the hydraulicfluid that is fed by the second hydraulic fluid inlet 175 at itsperipheral surface. The overall length of the third plunger 183 is equalto the depth of the third plunger hole 203. The end of the third plunger183 that abuts the fourth plunger 184 is formed in spherical shape.

FIG. 36(a) shows the condition in which the first rocker arm 162 isdetached from the second rocker arm 164 and third rocker arm 166 andhydraulic fluid pressure is supplied to the first passage 172 forsupply/discharge of hydraulic fluid (hereinbelow this is termed“hydraulic pressure ON”), but hydraulic fluid pressure is not suppliedto the second passage 174 for supply/discharge of hydraulic fluid(hereinbelow this is termed “hydraulic pressure OFF”). By turning thehydraulic pressure of the first hydraulic fluid inlet 173 that is fedfrom the first passage 172 for supply/discharge of hydraulic fluid ON,the first plunger 181 is pressed rightwards (direction of the arrow inthe Figure). This pressing force is larger than the biasing force of thesecond plunger spring 187, so the first plunger 181 is moved rightwardsintegrally with the second plunger 182. Since the overall length of thesecond plunger 182 is equal to the depth of the second plunger hole 202,the contact point of the first plunger 181 and the second plunger 182 ison the mating plane of the first rocker arm 162 and second rocker arm164.

On the other hand, the hydraulic pressure of the second hydraulic fluidinlet 175 that is fed from the second passage 174 for supply/dischargeof hydraulic fluid is OFF, so the third plunger 183, fourth plunger 184and fifth plunger 185 are moved leftwardly (direction of the arrow inthe Figure) integrally, by the biasing force of the fifth plunger spring189. Since the overall length of the third plunger 18.3 is equal to thedepth of the third plunger hole 203, the contact point of the thirdplunger 183 and the fourth plunger 184 is on the mating plane of thefirst rocker arm 162 and third rocker arm 166. Furthermore, since theoverall length of the fourth plunger 184 is equal to the depth of thefourth plunger hole 204, the contact point of the fourth plunger 184 andthe fifth plunger 185 is on the mating plane of the first rocker arm 162and of the second rocker arm 164.

Thus, since the contact point of the plungers is on the mating plane ofthe rocker arms, the first rocker arm 162 is in a condition detachedfrom the second rocker arm 164 and third rocker arm 166. The firstrocker arm 162 therefore performs operation in accordance with the firstcam 152 that abuts the upper surface thereof i.e. rocking about therocker shaft 170 is deactivated, as a result of which the burnt gasintroduction valve 31 a is deactivated in the closed condition.

FIG. 36(b) shows the condition in which the first rocker arm 162 islinked only with the second rocker arm 164 and both the first passage172 for supply/discharge of hydraulic fluid and the second passage 174for supply/discharge of hydraulic fluid are in the hydraulic pressureOFF condition. Since the first hydraulic fluid inlet 173 that is fedfrom the first passage 172 for supply/discharge of hydraulic fluid is inthe hydraulic pressure OFF condition, the first plunger 181 and thesecond plunger 182 are moved leftwardly (direction of the arrow in theFigure) by the biasing force of the second plungers spring 187. Sincethe overall length of the first plunger 181 is shorter than the depth ofthe first plunger hole 201, part of the second plunger 182 enters thefirst plunger hole 201.

Furthermore, since the second hydraulic fluid inlet 175 that is fed fromthe second passage 174 for supply/discharge of hydraulic fluid is in thehydraulic pressure OFF condition, in the same way as in the case of FIG.36(a), the contact point of the 180 third plunger 183 and the fourthplunger 184 is in the mating plane of the first rocker arm 162 and thirdrocker arm 166, while the contact point of the fourth plunger 184 andfifth plunger 185 is in the mating plane of the first rocker arm 162 andsecond rocker arm 164.

In this way, the first rocker arm 162 is linked with the second rockerarm 164 by entry of part of the second plunger 182 into the firstplunger hole 201. Also, since the contact point of the third plunger 183and the fourth plunger 184 is in the mating plane of the first rockerarm. 162 and third rocker arm 166, the first rocker arm 162 and thethird rocker arm 166 are put in detached condition. The first rocker arm162 therefore performs operation under the control of the second cam 154that abuts the upper surface of the second rocker arm 164. That is,rocking takes place downwardly by a prescribed amount at a prescribedcrank angle as the camshaft 151 rotates, thereby opening and closing theburnt gas introduction valve 31 b.

FIG. 36(c) shows the condition in which the first rocker arm 162 islinked with the second rocker arm 164 and third rocker arm 166; thefirst passage 172 for supply/discharge of hydraulic fluid is in thehydraulic pressure OFF condition and the second passage 174 forsupply/discharge of hydraulic fluid is in the hydraulic pressure ONcondition. Since the hydraulic pressure of the first hydraulic fluidinlet 173 that is fed from the first passage 172 for supply/discharge ofhydraulic fluid is OFF, in the same way as in FIG. 36(b), part of thesecond plunger 182 enters the first plunger hole 201.

Furthermore, since the hydraulic pressure of the second hydraulic fluidinlet 175 that is fed from the second passage 174 for supply/dischargeof hydraulic fluid is ON, the third plunger 183 is pressed rightwardly(direction of the arrow in the Figure). The pressing force thereof isgreater than the biasing force of the fifth plunger spring 189, so thethird plunger 183 is moved rightwardly integrally with the fourthplunger 184 and fifth plunger 185. Since the overall length of the fifthplunger 185 is shorter than the depth of the fifth plunger hole 205,part of the fourth plunger 184 enters the fifth plunger hole 205 and, inaddition, part of the third plunger 183 enters the fourth plunger hole204.

Thus, since part of the second plunger 182 enters the first plunger hole201 and part of the third plunger 183 enters the fourth plunger hole204, the first rocker arm 162 is linked with the second rocker arm 164and of the third rocker arm 166. Consequently, the first rocker arm 162performs operation under the control of the third cam 156, which is ofthe largest radius of rotation and abuts. the upper surface of the thirdrocker arm 166. That is, rocking takes place downwards by a prescribedamount at a prescribed crank angle as the camshaft 151 is rotated,thereby opening and closing the burnt gas introduction valve 31 b. Theperiod of this valve opening is longer than in the case of FIG. 36(b).

The above cam changeover mechanism 150 was provided for the burnt gasintroduction valves 31 b and burnt gas exhaust valves 32 b, but asimilar cam changeover mechanism 150 a (shown in brackets in FIG. 35) isprovided in respect of the following cylinders' intake valves 31 a andpreceding cylinders' exhaust valves 32 a. However, in the case of thecam changeover mechanism 150 a, the second cam 154 and third cam 156 areof the same shape. The rocker arm set 160 a that abuts these camscomprises a first rocker arm 162 a, second rocker arm 164 a and thirdrocker arm 166 a, as shown in FIG. 35. The cam changeover mechanism 150a effects changeover between a condition in which a deactivatedcondition is produced by detaching the first rocker arm 162 a from thesecond rocker arm 164 a and third rocker arm 166 a, putting thefollowing cylinders' intake valves 31 a and preceding cylinders' exhaustvalves 32 a in the closed condition, and a condition in which the firstrocker arm 162 a is linked with the second rocker arm 164 a and thirdrocker arm 166 a, so that the following cylinders' intake valves 31 aand preceding cylinders' exhaust valves 32 a are opened and closed byrotation of the second cam 154 and third cam 156.

FIG. 37 is a diagram illustrating the operation of the three plungersprovided within the rocker arm set 160 a. FIG. 37(a) shows the conditionin which the first rocker arm 162 a is detached from the second rockerarm 164 a and third rocker arm 166 a and FIG. 37(b) shows the conditionin which the first rocker arm 162 a is linked with the second rocker arm164 a and third rocker arm 166 a.

The plunger construction in the rocker arm set 160 a is that provided bythe construction represented by the third plunger 183, fourth plunger184 and fifth plunger 185 of the rocker arm set 160 so a description ofthe detailed construction thereof is duplicated in the description ofthe rocker arm set 160 and so will be dispensed with. However, itdiffers from the rocker arm set 160 in that a first hydraulic fluidinlet 173 a leads from a first passage 172 for supply/discharge ofhydraulic fluid at the left end of the third plunger 183. Also, thesecond passage 174 for supply/discharge of hydraulic fluid may bedispensed with if it can be omitted structurally.

FIG. 37(a) shows the condition in which the first rocker arm 162 isdetached from the second rocker arm 164 and third rocker arm 166 and thefirst passage 172 for supply/discharge of hydraulic fluid is in thehydraulic pressure OFF condition. Since the first hydraulic fluid inlet173 a that is fed from the first passage 172 for supply/discharge ofhydraulic fluid is in the hydraulic pressure OFF condition, the thirdplunger 183, fourth plunger 184 and fifth plunger 185 are movedleftwardly (direction of the arrow in the Figure) integrally by thebiasing force of the third plunger spring 189. Consequently, the contactpoint of the third plunger 183 and the fourth plunger 184 is in themating plane of the first rocker arm 162 a and third rocker arm 166 aand the contact point of the fourth plunger 184 and fifth plunger 185 isin the mating plane of the first rocker arm 162 a and second rocker arm174 a.

Thus, since the contact points of the plungers are in the mating planesof the rocker arms, the first rocker arm 162 a is detached from thesecond rocker arm 164 a and third rocker arm 166 a. Consequently, thefirst rocker arm 162 a performs operation under the control of the firstcam 152 abutting the upper surface thereof i.e. rocking about the rockershaft 170 is deactivated, so the burnt gas introduction valve followingcylinders' intake valve 31 a and the preceding cylinders' exhaust valve32 a are deactivated in the closed condition.

FIG. 37(b) shows the condition in which the first rocker arm 162 a islinked with the second rocker arm 164 a and third rocker arm 166 a andthe first passage 172 for supply/discharge of hydraulic fluid is in thehydraulic pressure ON condition. Since the first hydraulic fluid inlet173 a that is fed from the first passage 172 for supply/discharge ofhydraulic fluid is in the hydraulic pressure ON condition, the thirdplunger 183 is pressed rightwardly (direction of the arrow in theFigure). Since the pressing force thereof is greater than the biasingforce of the fifth plunger spring 189, the third plunger 183 is movedrightwardly integrally with the fourth plunger 184 and fifth plunger185. Consequently, part of the fourth plunger 184 enters the fifthplunger hole 205 and part of the third plunger 183 enters the fourthplunger hole 204.

Thus, since part of the fourth plunger 184 enters the fifth plunger hole205 and a part of the third plunger 183 enters the fourth plunger hole204, the first rocker arm 162 is linked with the second rocker arm 164and the third rocker arm 166. The first rocker arm 162 is thereforeoperated under the control of the second cam 154 and third cam 156 (ofthe same shape) that abut the upper surface of the second rocker arm 164a and third rocker arm 166 a. Specifically, rocking takes placedownwardly by a prescribed amount at a prescribed crank angle withrotation of the camshaft 151, causing the burnt gas introduction valve31 b to be opened and closed.

FIG. 38 shows the construction of the drive and control systems when thecam changeover mechanism shown in FIG. 35 to FIG. 37 is employed. Afirst control valve 176 and second control valve 177 are control valvesfor control of (hydraulic pressure ON/OFF) of first hydraulic fluid andsecond hydraulic fluid fed to the first passage 172 for supply/dischargeof hydraulic fluid and second passage 174 for supply/discharge ofhydraulic fluid. A cam changeover mechanism 150 is provided at the burntgas introduction valve 31 a and burnt gas exhaust valve 32 b; a camchangeover mechanism 150 a is provided at the following cylinders'intake valve 31 a and preceding cylinders' exhaust valve 32 a.

The ECU 70 a differs from the ECU70 shown in FIG. 31 in that instead ofthe valve deactivating mechanism controller 72 and cam phase controller79 it comprises cam changeover controller 190.

The cam changeover controller 190 controls the cam changeover mechanism150 and cam changeover mechanism 158 as follows by controlling the firstcontrol valve 176 and second control valve 177 in accordance with thespecial operating mode and ordinary operating mode or in accordance withthe operating condition.

In the special operating mode, the low load, low speed region is:

-   -   First hydraulic pressure-OFF and second hydraulic pressure-OFF    -   Preceding cylinders' exhaust valves 32 a and following        cylinders' intake valves 31 a in deactivated condition (FIG.        37(a))    -   Burnt gas exhaust valves 32 b and burnt gas introduction valves        31 b in operating condition under the control of the second cam        154 (low speed cam) (FIG. 36(b));

In the special operating mode, the high load, high speed region is:

-   -   First hydraulic pressure-OFF and second hydraulic pressure-ON    -   Preceding cylinders' exhaust valves 32 a and following        cylinders' intake valves 31 a in deactivated condition (FIG.        37(a))    -   Burnt gas exhaust valves 32 b and burnt gas introduction valves        31 b in operating condition under the control of the third cam        156 (high speed cam) (FIG. 36(c));

The ordinary operating mode is:

-   -   First hydraulic pressure-ON and second hydraulic pressure-OFF    -   Preceding cylinders' exhaust valves 32 a and following        cylinders' intake valves 31 a in operating condition under the        control of the second cam 154 and third cam 156 (37(b))    -   Burnt gas exhaust valves 32 b and burnt gas introduction valves        31 b in deactivated condition (FIG. 36(a)).

Next, the operation of the device shown in FIG. 35 to FIG. 38 will bedescribed; however, description of portions that have already beendescribed with reference to the embodiment shown in FIG. 31 and FIG. 32will be dispensed with. FIG. 39 is a diagram showing in detail theportions associated with the intake/exhaust strokes in this embodiment.FIG. 39(a) is the case of the comparatively low load, low rotationalspeed region (region A401 in FIG. 32) of an operating condition in whichthe special operating mode is being performed and FIG. 39(b) is the caseof the comparatively high load, high rotational speed region (regionA402 in FIG. 32) of this mode. Other notation is the same as in FIG. 33.

The top section of FIG. 39(a) shows the open period 230 (shown shaded)of the burnt gas exhaust valves for which the burnt gas exhaust valves32 b of the preceding cylinders 2A, 2D are open and the open period 231of the preceding cylinders' intake valves for which the precedingcylinders' intake valves 31 are open. The bottom section shows the openperiod 232 of the following cylinders' exhaust valves for which thefollowing cylinders' exhaust valves 32 of the following cylinders 2B, 2Care open and the open period 233 (shown shaded) of the burnt gasintroduction valves. It also, directly below the open period 230 of theburnt gas exhaust valves and the open period 233 of the burnt gasintroduction valves, the open period 250 of the burnt gas exhaust valvesand open period 253 of the burnt gas introduction valves, to bedescribed (both of these are for the case of comparatively highrotational speed, high load in the special operating mode) are shown forreference purposes by the double-dotted chain lines. Changeover of theseopen periods is performed by means of the cam changeover mechanisms 150,150 a and the cam changeover controller 190. The open period 231 of thepreceding cylinders' intake valves is set from about 10° CA before TDCto about 55° CA after BDC (total about 245° CA). This is the typical setvalue of a prior art engine. In contrast, the open period 233 of theburnt gas introduction valves is set from substantially the TDC to theBDC (total about 180° CA). Also, the open period 232 of the followingcylinders' exhaust valves is set from about 45° CA before BDC to about10° CA after TDC (total about 235° CA).

That is, the interval (about 0° CA) between the following cylinders'intake stroke and bottom dead center 247 and the closure time 248 of theburnt gas introduction valves is set to be shorter than the interval(about 55° CA) between the preceding cylinders' intake stroke bottomdead center 242 and the closure time 243 of the preceding cylinders'intake valves. The open period 233 of the burnt gas introduction valvesis shorter than the open period 231 of the preceding cylinders' intakevalves and the open period 232 of the following cylinders' exhaustvalves. Thus, since, in the following cylinders 2B, 2C, the open period233 of the burnt gas introduction valves is set to be short and theopening time 248 of the burnt gas introduction valves is set early, at atime close to the BDC, even during rise of the piston 3 after the BDC,either there is no period for which the burnt gas introduction valves 31b are open, or such a period is short. The effective compression ratiois therefore increased and is close to the geometrical compressionratio. With this increase in the effective compression ratio, thecylinder temperature of the following cylinders 2B, 2C tends to rise,increasing the compression ignition capability.

Also, the open period 230 of the burnt gas exhaust valves is set fromabout 45° CA before the BDC to about 20° CA before the TDC (total about205° CA) Consequently, the opening time 241 of the burnt gas exhaustvalves is set advanced by about 20° CA from the closure time 248 of theburnt gas introduction valves (corresponding to the TDC in the precedingcylinders 2A, 2D). This is earlier than the typical set value in aconventional engine (about 50° CA after the TDC). By closing the burntgas exhaust valve 32 b earlier, the internal EGR of the precedingcylinders 2A, 2D is increased, so the temperature of the burnt gas thatis introduced into the following cylinders 2B, 2C rises.

Since the open period 232 of the following cylinders' exhaust valves isset from about 45° CA before the BDC to about 10° CA after the TDC(total about 235° CA), the following cylinders' exhaust valves 32 areopen at the top dead center 244 of the following cylinders' exhauststroke. On the other hand, the opening time 246 of the burnt gasintroduction valve is set so as to be practically at the top dead center245 of the following cylinders' intake stroke. That is, the opening time246 of the burnt gas introduction valves is set to be at the top deadcenter 245 of the following cylinders' intake stroke, whilst thefollowing cylinders' exhaust valves 32 are open until the followingcylinders' exhaust stroke top dead center 244. Consequently, since thevalve overlap in the vicinity of the top dead center 244 of thefollowing cylinders' exhaust stroke is shortened, so-called “blowthrough” in which the burnt gas that is introduced into the followingcylinders 2B, 2C is directly exhausted to the exhaust passages 20through the following cylinders' exhaust valves 32 can be prevented andthe effective compression ratio of the following cylinders 2B, 2C can beincreased.

As described above, the cylinder temperature of the following cylinders2B, 2C is increased and the compression ignition capability improved byincreasing the effective compression ratio of the following cylinders2B, 2C and raising the temperature of the burnt gas that is introducedinto the following cylinders 2B, 2C. In this way, the operating regionin which combustion by compression ignition can be performed is expandedfurther into the low load region.

FIG. 39(b) is a diagram of the case of the region of comparatively highload and high rotational speed (region A402 in FIG. 32). The uppersection thereof shows the open period 250 (shown cross-hatched) of theburnt gas exhaust valves, for which the burnt gas exhaust valves 32 b ofthe preceding cylinders 2A, 2D are open and the open period 231 (commonwith FIG. 39(a)) of the preceding. cylinders' intake valves. The lowersection thereof shows the open period 232 (common with FIG. 39(a)) ofthe following cylinders' exhaust valves and the open period 253 (showncross-hatched) of the burnt gas introduction valves, in which the burntgas introduction valves 31 b are open. Also, directly below the openperiod 250 of the burnt gas exhaust valves and the open period 253 ofthe burnt gas introduction valves, the open period 230 of the burnt gasexhaust valves and the open period 233 of the burnt gas introductionvalves of FIG. 39(a) are shown with double-dotted chain lines forreference.

The open period 253 of the burnt gas introduction valves is set fromabout 10° CA before the TDC to about 55° CA after the BDC. That is,valve opening takes place 10° CA earlier than the open period 233 of theburnt gas introduction valves in the low load, low rotational speedregion and valve closure takes place 55° later. Consequently, theeffective compression ratio of the following cylinders 2B, 2C is reducedcompared with the case of FIG. 39(a) and the cylinder temperature of thefollowing cylinders 2B, 2C is lowered.

Also, the open period 250 of the burnt gas exhaust valves is set fromabout 45° CA before the BDC to about 10° CA after the TDC. That is, itis about 30° CA longer than the open period 230 of the burnt gas exhaustvalves of FIG. 39(a). Internal EGR of the preceding cylinders 2A, 2D istherefore reduced compared with FIG. 39(a), so the temperature of theburnt gas that is introduced into the following cylinders 2B, 2C falls.

As described above, the cylinder temperature of the following cylinders2B, 2C is lowered and abnormal combustion such as knocking is preventedby decreasing the effective compression ratio of the following cylinders2B, 2C and lowering the temperature of the burnt gas that is introducedinto the following cylinders 2B, 2C. In this way, the operating regionin which combustion by compression ignition can be performed is furtherexpanded into the high load region.

It should be noted that the pattern shown in FIGS. 39(a) and (b) is notrestrictive of the opening periods and closure periods and these may besuitably altered within the scope of the claims. For example, theclosure time 248 of the burnt gas introduction valves in FIG. 39(a) maybe set to be somewhat later than the bottom dead center 247 of thefollowing cylinders' intake stroke. Other specific set values may be setto suitable values in accordance with the demanded engine performance.

In these embodiments, constructions were adopted in which an arrangementwith three-way cam changeover, an arrangement with two-way camchangeover and an arrangement with no cam changeover were combined inaccordance with the location of arrangement of the intake/exhaust valvesand their respective functions, but the combinations thereof are notrestricted to these; for example, three-way cam changeover could beemployed by applying three-way cams and rocker arm sets 160 to all ofthe valves. Also, the open period 231 of the preceding cylinders' intakevalves and the open period 232 of the following cylinders' exhaustvalves, which were taken as fixed in the second embodiment, could bevaried by cam changeover.

FIG. 40 to FIG. 44 show yet a further embodiment of the presentinvention.

In the overall construction of the engine shown in FIG. 40, the mainengine body, intake/exhaust ports, intake/exhaust passages connectedthereto and the intake/exhaust valves that open and close theinter-cylinder gas passages and intake/exhaust ports are constructed inthe same way as in the case of the embodiments shown in FIG. 30. Inaddition, the cam phase varying mechanisms 33 a and 34 a that areprovided in respect of the camshafts 33 and 34 are constructed in thesame way as in the case of the embodiment shown in FIG. 30. Upstream ofthe merging section in the intake passage 15, there are provided anairflow sensor 19 that detects the amount of the intake current, acompressor 27 b of a turbo supercharger 27 that performs superchargingutilizing the energy of the exhaust gas, and an inter-cooler 28. Also,in the exhaust passage 10, there are provided an O₂ sensor 23, turbine27 a of the turbo supercharger 27 and a three-way catalyst 24.

Also, as shown in FIG. 42, in respect of the burnt gas introductionvalves 31 b etc, there are provided cam changeover mechanisms 150 thatchange the rocking condition of the rocker arm sets 160 by camchangeover; by means of these mechanisms, it is arranged to be possibleto vary the opening/closing times of the burnt gas introduction valves31 b etc shown by the double-dotted chain lines or to deactivate thesein the closed condition.

A camshaft 34 is arranged above the burnt gas introduction valves 31 b.Cams 27 are arranged so as to rotate integrally with this camshaft 34.The cams 27 comprise three cams having independent lift characteristics,namely, a first cam 152, second cam 154 and third cam 156. A rocker armset 160 that is supported by a rocker shaft 170 is provided betweenthese cams and the burnt gas introduction valves 31 b. This rocker armset 160 is of the same construction as that illustrated in FIG. 36,described above.

The cam changeover mechanism 150 is provided for the burnt gasintroduction valves 31 b and burnt gas exhaust valves 32 b, but asimilar cam changeover mechanism 150 a and rocker arm set 160 a (shownin brackets in FIG. 42) are also provided for the following cylinders'intake valves 31 a and preceding cylinders' exhaust valves 32 a.However, in the cam changeover mechanism 150 a, the second cam 154 andthird cam 156 are of the same shape. The rocker arm set 160 a that abutsthese cams is of the same construction as that illustrated in FIG. 35and FIG. 37 described above.

In addition, a similar cam changeover mechanism 150 b is also providedfor the preceding cylinders' intake valves 31 (shown in brackets in FIG.42). In the cam changeover mechanism 150 b, the first cam 152 is not adeactivation cam but is a cam for low load, having a projection. Also,the second cam 154 and third cam 156 are of the same shape, being camsfor high load. The rocker arm set 160 b that abuts these cams comprisesa first rocker arm 162 b, second rocker arm 164 b and third rocker arm166 b, as shown in brackets in FIG. 42. The cam changeover mechanism 150b changes over between a condition in which the preceding cylinders'intake valves 31 are opened for a comparatively short period bydetachment of the first rocker arm 162 b from the second rocker arm 164b and third rocker arm 166 b and a condition in which the first rockerarm 162 b is linked with the second rocker arm 164 b or third rocker arm166 b, so that the preceding cylinders' intake valves 31 are opened fora comparatively long period by rotation of the second cam 154 or thirdcam 156.

FIG. 43 is a diagram showing the operation of the three plungersprovided in the interior of the rocker arm cassette 160 b. FIG. 43(a)shows the condition in which the first rocker arm 162 b is detached fromthe second rocker arm 164 b and the third rocker arm 166 b; FIG. 43(b)shows the condition in which the first rocker arm 162 b is linked withthe second rocker arm 164 b only; and FIG. 43(c) shows the condition inwhich the first rocker arm 162 b is linked with the second rocker arm164 b and the third rocker arm 166 b.

In the plunger construction in the rocker arm set 160, the plungerconstruction in the rocker arm set 160 b is altered in respect of theoverall length of the first plunger 181 and second plunger 182 and thespecification of the second plunger spring 187; these will thereforerespectively be referred to as the first plunger 181 b, second plunger182 b and second plunger spring 187 b. Due to this construction, the camchangeover condition produced by the cam changeover mechanism 150 b bythe combination of the hydraulic pressure of the first passage 172 forsupply/discharge of hydraulic fluid and the second passage 174 forsupply/discharge of hydraulic fluid being turned ON and OFF differs fromthat of the cam changeover mechanism 150.

FIG. 43(a) shows the condition in which the first rocker arm 162 b isdetached from the second rocker arm 164 b and the third rocker arm 166 band the first passage 172 for supply/discharge of hydraulic fluid andthe second passage 174 for supply/discharge of hydraulic fluid are bothin the hydraulic pressure OFF condition. FIG. 43(b) shows the conditionin which the first rocker arm 162 b is linked with the second rocker arm164 b only and the first passage 172 for supply/discharge of hydraulicfluid is in the hydraulic pressure ON condition while the second passage174 for supply/discharge of hydraulic fluid is in the hydraulic pressureOFF condition. FIG. 43(c) shows the condition in which the first rockerarm 162 b is linked with the second rocker arm 164 b and the thirdrocker arm 166 b and the first passage 172 for supply/discharge ofhydraulic fluid is in the hydraulic pressure OFF condition while thesecond passage 174 for supply/discharge of hydraulic fluid is in thehydraulic pressure ON condition.

FIG. 44 shows the construction of the drive and control systems in thisembodiment; in this Figure, the ECU 70 b comprises a operating conditionidentifier 71, cam changeover controller 190, intake air quantitycontroller 73, combustion controller 74 and cam phase controller 77.

By examining the engine operating condition (engine rotational speed andengine load) using the signals from the engine speed sensor 77 andaccelerator pedal stroke sensor 78 etc, the operating conditionidentifier 71 ascertains whether the operating condition is in theoperating region A on the low load, low rotational speed side (engineload less than T1 and engine rotational speed less than r1) as shown inFIG. 45 or is in the operating region B on the high load or highrotational speed side (engine load greater than T1 or engine rotationspeed greater than r1). In the operating region A, the operating regionA501 is a region of comparatively low load, low rotational speed, theoperating region A503 is a region of comparatively high load, highrotational speed and operating region A502 is an intermediate regionthereof. Under prescribed conditions (for example condition in which theengine is fully warmed-up), operation is performed in the specialoperating mode, in which the cylinders are put in two-cylinder connectedcondition, in the operating region A and operation is conducted in theordinary operating mode in which the respective cylinders are inindependent condition, in the operating region B.

The cam changeover controller 190 controls the cam changeover mechanism150 and cam changeover mechanism 150 a as follows by controlling thefirst control valve 176 and second control valve 177 in accordance withwhether the engine is in the special operating mode or the ordinaryoperating mode or in accordance with the operating region.

In the special operating mode, the low/intermediate regions (regionsA501, A502) are:

-   -   First hydraulic pressure-OFF and second hydraulic pressure-OFF    -   Preceding cylinders' exhaust valves 32 a and following        cylinders' intake valves 31 a in deactivated condition (FIG.        37(a)) under the control of the first cam 152;    -   Burnt gas exhaust valves 32 b and burnt gas introduction valves        31 b in operating condition under the control of the second cam        154 (low load cam) (FIG. 36(b));    -   Preceding cylinders' intake valves 31 in operating condition        (FIG. 43(a)) under the control of the first cam 152 (low load        cam);

In the special operating mode, the high load region (region A503) is:

-   -   First hydraulic pressure-OFF and second hydraulic pressure-ON    -   Preceding cylinders' exhaust valves 32 a and following        cylinders' intake valves 31 a in deactivated condition (FIG.        37(a)) under the control of the first cam 152;    -   Burnt gas exhaust valves 32 b and burnt gas introduction valves        31 b in operating condition under the control of the third cam        156 (high load cam) (FIG. 36(c));    -   Preceding cylinders' intake valves 31 in operating condition        (FIG. 43(b)) under the control of the second cam 154 b (high        load cam);

The ordinary operating mode (region B) is:

-   -   First hydraulic pressure-ON and second hydraulic pressure-OFF    -   Preceding cylinders' exhaust valves 32 a and following        cylinders' intake valves 31 a in operating condition under the        control of the second cam 154 and third cam 156 (37(b));    -   Burnt gas exhaust valves 32 b and burnt gas introduction valves        31 b in deactivated condition (FIG. 36(a)) under the control of        the first cam 152;    -   Preceding cylinders' intake valves 31 in operating condition        (FIG. 43(c)) under the control of the third cam 156 b (high load        cam).

The air intake rate controller 73 controls the air intake rate in thesame way as the air intake rate controller 73 in FIG. 38.

The combustion controller 74 comprises fuel injection controller 75 andignition controller 76. In the low/intermediate load region of thespecial operating mode (region A501 and A502 of FIG. 45), in respect ofthe preceding cylinders (first and fourth cylinders 2A, 2D), it controlsthe fuel injection rate such that the air/fuel ratio is a lean air/fuelratio greater than the stoichiometric air/fuel ratio, preferablysubstantially twice or more the stoichiometric air/fuel ratio, and, inrespect of the following cylinders (second and third cylinders 2B, 2C),it supplies fuel in respect of the burnt gas of lean air/fuel ratio thatis introduced from the preceding cylinders and controls the fuelinjection rate such that the air/fuel ratio is an air/fuel ratio ofsubstantially the stoichiometric air/fuel ratio, or leaner than this.These fuel injections are performed in the intake stroke, so thatcombustion by compression self-ignition is performed in both thepreceding cylinders 2A, 2D and the following cylinders 2B, 2C.

Also, in the high load region of the special operating mode (region A503of FIG. 4), the fuel injection rate in respect of the precedingcylinders (first and fourth cylinders 2A, 2D) is controlled such thatthe air/fuel ratio is a lean air/fuel ratio larger than thestoichiometric air/fuel ratio, preferably substantially twice or morethe stoichiometric air/fuel ratio and the ignition timing is set such asto produce forced ignition in the vicinity of the compression top deadcenter. In respect of the following cylinders (second and thirdcylinders 2B, 2C), fuel is supplied in respect of the burnt gas of leanair/fuel ratio that is introduced from the preceding cylinders, and thefuel injection rate is controlled and the injection timing with whichfuel is injected in the intake stroke is set such that the air/fuelratio is substantially the stoichiometric air/fuel ratio. Combustion isperformed by compression self-ignition or forced ignition, in accordancewith the operating condition.

Control in the case of the ordinary operating mode is the same as in thecase of the other embodiments described above.

Also, the cam phase controller 77 controls the cam phase varyingmechanisms 33 a, 34 a in accordance with the results of the operatingcondition identifier 71. The details of the control action are describedlater, but for example in the special operating mode, in the low loadregion (region A501 of FIG. 45), the cam phase varying mechanism 33 a iscontrolled so as to delay the phase of the cam 26 and the cam phasevarying mechanism 34 a is controlled so as to advance the phase of thecam 27. There is therefore an overall delay in the opening/closure timesof the preceding cylinders' intake valves 31 a and following cylinders'intake valves 31 a that are operated by rotation of the camshaft 33 andan overall advancement of the opening/closure times of the burnt gasexhaust valves 32 b, burnt gas introduction valves 31 b and followingcylinders' exhaust valves 32 that are operated by rotation of the camshaft 34. In contrast, in the medium/high load region (regions A502,A503 of FIG. 45) or the ordinary operating mode region (region B of FIG.45), the phases of the cams 26, 27 are controlled respectively in theopposite direction, with the result that the opening/closure times ofthe preceding cylinders' intake valves 31 and following cylinders'intake valves 31 a are, overall, advanced, and that the opening/closuretimes of the burnt gas discharge valves 32 b, burnt gas introductionvalves 31 b and following cylinders' exhaust valves 32 are, overall,delayed. It should be noted that, since the cam phase varying mechanisms33 a, 34 a operate on the valves during operation thereof, valves whichare in a deactivated condition are maintained in this deactivatedcondition irrespective of control by the cam phase varying mechanisms 33a, 34 a.

Next, the operation of a device according to this embodiment will bedescribed with reference to FIG. 46 to FIG. 48.

In a device according to this embodiment, in the operating region A oflow load, low rotational speed, the device put into the specialoperating mode in which combustion is conducted in a condition with twocylinders connected and in the high load or high rotational speedoperating region the device is put in the ordinary operating mode inwhich combustion is conducted in a condition with the intake ports andexhaust ports of the respective cylinders independent. Also, when in thespecial operating mode, combustion is conducted with a super-leanair/fuel ratio in the preceding cylinders and combustion is conducted bycompression self-ignition in the following cylinders.

That is, even in the operating region in the special operating mode inwhich compression self-ignition is performed, when in a condition inwhich the cylinder temperature is low and compression self-ignition isdifficult, all of the cylinders are changed over to forced ignition.Also, contrariwise, if the cylinder temperature of the followingcylinders 2B, 2C becomes too high, producing a condition in whichabnormal combustion such as knocking tends to occur the device ischanged over from the special operating mode to the ordinary operatingmode. In both cases, the benefit of improved fuel costs etc issuppressed compared with the special operating mode using compressionself-ignition. Consequently, in order to obtain these benefits to agreater extent, it is desirable to expand the operating region in thespecial operating mode that is suitable for compression self-ignitionand, in addition, it is desirable to expand the operating region in thepreceding cylinders 2A, 2D that is suitable for compressionself-ignition.

The opening/closure times of the intake/exhaust valves are set asfollows in order to expand the operation region that is suitable forcompression self-ignition in compression self-ignition in the specialoperating mode, in particular in the preceding cylinders 2A, 2D.

FIG. 46 is a diagram showing the opening/closure times of the precedingcylinders' intake valves 31 and burnt gas exhaust valves 32 b of thepreceding cylinders 2A, 2D in the special operating mode and theopening/closure times of the burnt gas introduction valves 31 b of thefollowing cylinders 2B, 2C and the following cylinders' exhaust valves32, showing the intake/exhaust stroke portion in detail. FIG. 46(a) isthe case of the comparatively low load region (region A501 of FIG. 45)and FIG. 46(b) is the case of the intermediate load region (region A502of FIG. 45), of the operating region in which the special operating modeis performed. In these Figures, the horizontal axis shows the crankangle, T is the top dead center (TDC) and B is the bottom dead center(BDC). The interval between DNB is 180° CA. Also, the upper sectionshows the preceding cylinders 2A, 2D and the lower section shows thefollowing cylinders 2B, 2C corresponding to these. Also, the portionsindicated by band lines show the open periods of the respective valves.The white arrows from the upper section to the lower section indicatethe condition in which the exhaust strokes of the preceding cylinders2A, 2D and the intake strokes of the following cylinders 2B, 2C overlapand gas burnt in the preceding cylinders 2A, 2D is conducted into thefollowing cylinders 2B, 2C.

FIG. 46(a) shows in the upper section thereof the open period 300 of theburnt gas exhaust valves, for which the burnt gas exhaust valves 32 b ofthe preceding cylinders 2A, 2D are open and the open period 310 (shownshaded) of the preceding cylinders' intake valves, for which thepreceding cylinders' intake valves 31 are open. Both of these openperiods are produced by the low load cam. Below the respective bandlines, the open periods produced by the high load cam (open period 380of the burnt gas exhaust valves and open period 390 of the precedingcylinders' intake valves) are shown for reference. The bottom sectionshows the open period 320 of the following cylinders' exhaust valves forwhich the following cylinders' exhaust valves 32 of the followingcylinders 2B, 2C are open and the open period 330 of the burnt gasintroduction valves, for which the burnt gas introduction valves 31 bare open. The open period 300 of the burnt gas exhaust valves of thepreceding cylinders is set from about 35° CA before the BDC to about 40°CA before the TDC (total about 175° CA). In particular, the closure time302 of the burnt gas exhaust valves is set to be earlier than the TDC.This is a setting that is shorter and earlier than the generally setvalues for a conventional engine (from 30° CA before the BDC to about25° CA after the TDC). Also, the open period 310 of the precedingcylinders' intake valves is set from about 50° CA after the TDC to about45° CA after the BDC (total about 175° CA). This is a setting that isshorter and later than the generally set values of a conventional engine(from 10° CA before the TDC to about 55° CA after the BDC). By means ofthese settings, a condition is produced (hereinbelow called “minusoverlap”) in which all of the intake/exhaust valves are closed, from theclosure time 302 of the burnt gas exhaust valves to the open time 312 ofthe preceding cylinders' intake valves. In FIG. 46(a), the minus overlapof the preceding cylinders is about 90° CA, on both sides of the TDC.Also, in the preceding cylinders 2A, 2D, the fuel injection time 305 ofthe preceding cylinders is set in the vicinity of the TDC, so thatcombustion by compression self-ignition is performed.

The open period 320 of the following cylinders' exhaust valves and theopen period 330 of the burnt gas introduction valves are typical setvalues, but overall they are set to occur earlier. Also, the fuelinjection rates of the following cylinders 2B, 2C are set to be leanerthan the actual air/fuel ratio.

Thanks to the settings as above, in the preceding cylinders 2A, 2D,there is a large minus overlap, so there is a considerable amount ofinternal EGR. If there is a large amount of internal EGR, the shift tothe following intake stroke and compression stroke occurs in a conditionin which there is a large amount of burnt gas in the precedingcylinders, so the cylinder temperature is raised and compressionself-ignition is facilitated. Combustion in the preceding cylinders 2A,2D occurs by compression self-ignition, so a fuel consumptionimprovement effect and exhaust gas cleaning effect are produced by thehigh degree of thermal efficiency and suppression of generation of NOx.

If the amount of internal EGR of the preceding cylinders 2A, 2D isincreased in this way, on the other hand, the amount of new air in thefollowing cylinders 2B, 2C is liable to become insufficient. However, inthis embodiment, it is arranged to be possible to introduce sufficientoxygen into the following cylinders 2B, 2C even if there is aconsiderable amount of internal EGR in the preceding cylinders 2A, 2D,by increasing the absolute amount of new air (in particular oxygen), byperforming supercharging using the turbo supercharger 50. Also, theintake temperature is raised by the supercharging, so the capability forcompression self-ignition of the preceding cylinders 2A, 2D isincreased. Thus, by performing supercharging, the insufficiency of newair in the following cylinders 2B, 2C resulting from increase in theinternal EGR of the preceding cylinders 2A, 2D is mitigated and thecompression self-ignition capability in the preceding cylinders 2A, 2Dis improved, thereby expanding the operating range in which compressionself-ignition can be performed in the preceding cylinders.

Also, since the fuel injection time 305 of the preceding cylinders 2A,2D is set later than the closure time 302 of the burnt gas exhaustvalves and in the vicinity of the TDC of these cylinders, the fuel isinjected into the preceding cylinders 2A, 2D in which a considerableamount of burnt gas is still present. Activation of the injected fuel bythe high temperature is therefore achieved and, since the fuel isinjected early, in the vicinity of the TDC, activation is promoted to afully satisfactory extent and compression self-ignition capability isimproved. It should be noted that direct exhaustion of the injected fuelfrom the burnt gas exhaust valves 32 b is prevented by performing thefuel ignition after the open period 302 of the burnt gas exhaust valves.

Furthermore, since the air/fuel ratio even in the following cylinders2B, 2C is set to be leaner than substantially the stoichiometricair/fuel ratio, thermal efficiency is higher than if substantially thestoichiometric air/fuel ratio were set, so a large improvement in fuelcosts is obtained. Also, since generation of NOx is suppressed in boththe preceding cylinders 2A, 2D and the following cylinders 2B, 2C tovery great extent by the lean air/fuel ratio and the compressionself-ignition, exhaust gas cleansing performance is satisfied simply bythe provision of a three-way catalyst 24 (which may, if necessary, becombined with an oxygen catalyst) in the exhaust passage 20 a. In otherwords, a comparatively expensive lean NOx catalyst for reductiontreatment of the NOx is unnecessary, making it possible to lower costs.

FIG. 46(b) is the case of the intermediate load region (region A502 ofFIG. 45); the upper section thereof shows the open period 340 of theburnt gas exhaust valves and the open period 350 (shown shaded) of thepreceding cylinders' intake valves; the lower section thereof shows theopen period 360 of the following cylinders' exhaust valves and the openperiod 370 of the burnt gas introduction valves. Overall, the setting ofthe open period 340 of the burnt gas exhaust valves, the open period 360of the following cylinders' exhaust valves and the open period 370 ofthe burnt gas introduction valves is delayed by 30° CA in comparisonwith the open period 300 of the burnt gas exhaust valves, the openperiod 320 of the following cylinders' exhaust valves and the openperiod 330 of the burnt gas introduction valves of FIG. 46(a). This isachieved by delaying the face of the camshaft 34 by 30° CA by means ofthe cam phase varying mechanism 34 a. In contrast, the open period 350of the preceding cylinders' intake valves is set to advanced by 25° CAwith respect to the open period 310 of the preceding cylinders' intakevalves of FIG. 46(a). This is achieved by advancing the phase of thecamshaft 33 by 25° CA by means of the cam phase varying mechanism 33 a.Consequently, the closure time 342 of the burnt gas exhaust valves isabout 10° CA before the TDC and the opening time 352 of the precedingcylinders' intake valves is about 25° CA after the TDC. This minusoverlap is about 35° CA, which is 55° CA shorter than in the conditionof FIG. 46(a).

In this way, the minus overlap is shortened with increase in the load(from the operating region A1 to A2), so the internal EGR ratio of thepreceding cylinders 2A, 2D is decreased. Consequently, the ratio of newair is increased with increase in load, so ample new air for performingthe combustion is guaranteed in respect of the fuel injection rate,which has been increased due to the demanded output.

FIG. 47 is the case of the comparatively high load region (region A503of FIG. 45) of the operating condition in which the special operatingmode is performed. The notation is the same as in the case of FIG. 46(and also in the similar drawings below.). The upper section thereofshows the open period 380 of the burnt gas exhaust valves and the closedperiod 390 (shown shaded) of the preceding cylinders' intake valves.These are both open periods produced by the high load cam and arechanged over by the cam changeover mechanisms 150, 150 b from thecondition of FIG. 46. The open periods produced by the low load cam(open period 340 of the burnt gas exhaust valves and open period 350 ofthe preceding cylinders' intake valves are shown for reference below therespective and a lines. Also, the cam phase varying mechanism 33 a isset in the same way as in FIG. 46(b) and the cam phase varying mechanism34 a is set lagging the phase of the camshaft 34 by a further 5° CA fromthe condition of FIG. 46(b). As a result, the open period 380 of theburnt gas exhaust valves is set from about 30° CA before the BDC toabout 25° CA after the TDC (total about 235° CA) and the open period 390of the preceding cylinders' intake valves is set from about 10° CAbefore the TDC to about 55° CA after the BDC (total about 245° CA).Consequently, the closure period 382 of the burnt gas exhaust valves isset lagging by about 35° CA from the open period 392 of the precedingcylinders' intake valves. During this interval, both valves are open(hereinbelow this is termed “overlap”). These settings correspond to thetypical set values of conventional engine intake/exhaust valves. Also,in the preceding cylinders 2A, 2D, combustion is performed by forcedignition after making the air/fuel ratio leaner than in the case of theregions A1, A2. Fuel injection is changed over in the latter period ofthe compression stroke. Also, in the following cylinders 2B, 2C,combustion is performed by compression self-ignition in a condition inwhich the air/fuel ratio is substantially the stoichiometric air/fuelratio.

Thanks to such a setting, intake in the preceding cylinders 2A, 2D andexhaust of burnt gas are fully satisfactorily performed in thevalve-open period. In particular, since, in general, overlap is set, theinternal EGR amount is small, so there is an ample intake of new air, sothe required output can be obtained. The capability for compressionself-ignition decreases with decrease in the amount of internal EGR, butstable combustion is obtained by changing over to combustion by forcedignition. Thus a further improvement in regard to fuel costs isachieved, since the air/fuel ratio in the preceding cylinders 2A, 2D isset to be richer and the following cylinders 2B, 2C that are performingcombustion by compression self-ignition, and whose thermal efficiency iscorrespondingly improved, are set rather richer. Also, since thecombustion in the following cylinders is arranged to be performedsubstantially with the stoichiometric air/fuel ratio, fully satisfactoryexhaust gas cleansing performance can be obtained merely by theprovision of a three-way catalyst 24 arranged in the exhaust passage 20a.

FIG. 46 is a diagram showing the opening/closing times of the precedingcylinders' intake valves 31 and burnt gas exhaust valves 32 b of thepreceding cylinders 2A, 2D in the ordinary operating mode and theopening/closing times of the burnt gas introduction valves 31 b and thefollowing cylinders' exhaust valves 32 of the following cylinders 2B,2C. Since, in the ordinary operating mode, the cylinders are operatedindependently, combustion is performed by forced ignition by introducingnew air into both the preceding cylinders 2A, 2D and the followingcylinders 2B, 2C. The upper section shows the open period 420 of theburnt gas exhaust valves, for which the preceding cylinders' exhaustvalves 32 a of the preceding cylinders 2A, 2D are open and the openperiod 390 (shown shaded) of the preceding cylinders' intake valves, forwhich the preceding cylinders' intake valves 31 are open. The bottomsection shows the open period 440 of the following cylinders' exhaustvalves, for which the following cylinders' exhaust valves 32 of thefollowing cylinders 2B, 2C are open and the open period 450 of thefollowing cylinders' intake valves, for which the following cylinders'intake valves 31 a are open.

Exhaustion of the preceding cylinders 2A, 2D and intake in the followingcylinders 2B, 2C are performed by different valves than in the specialoperating mode, so their opening/closure is performed by different cams.Consequently, the open period 420 of the preceding cylinders' exhaustvalves and the open period 450 of the following cylinders' intake valvesare set independently of the open period of the burnt gas exhaust valvesand burnt gas introduction valves. In FIG. 48, the valves employed aredifferent, but the set values of the respective periods are the same asin the case of FIG. 47(b). Also, the open period 390 of the leadingvalves' intake valves and the open period 450 of the following valves'intake valves can be varied forwards and backwards by the cam phasevarying mechanism 33 a and the open period 420 of the precedingcylinders' exhaust valves and the open period 440 of the followingcylinders' exhaust valves can be varied forwards and backwards by meansof the cam phase varying mechanism 34 a. Overlap with the precedingcylinders can therefore be varied by controlling the cam phase varyingmechanisms 33 a, 34 a. The cam phase controller 49 is arranged such thatthe optimum thermal efficiency is obtained in accordance with the loadby controlling the cam phase varying mechanisms 33 a, 34 a such as toproduce a greater degree of valve overlap as the load is increased. Thesame control is performed in respect of the following cylinders 2B, 2C.

Thus, in the ordinary operating mode, output performance is ensured byexercising control to achieve the optimum intake/exhaust times,depending on the load, and controlling the air intake rates and fuelinjection rates such as to provide the stoichiometric air/fuel ratio ora ratio richer than this.

Next, a second example of control of intake/exhaust etc using a deviceaccording to this embodiment is described with reference to FIG. 49. Inthis example, the basic construction and basic control configuration arethe same as in the case of the first embodiment, but the phase of thecamshafts 33, 34 is varied by means of the cam phase varying mechanisms33 a, 34 a without changing over between a low load cam and high loadcam.

FIG. 49(a) shows the case of a comparatively low load region (regionA501 in FIG. 45) of the operating region in which special operating modeis being performed. The upper section thereof shows the open period 460of the burnt gas discharge valves and the open period 470 of thepreceding cylinders' intake valves; the lower section thereof shows theopen period 480 of the following cylinders' exhaust valves and the openperiod 490 of the burnt gas introduction valves. The open period 460 ofthe burnt gas exhaust valves is set from about 5° CA before the BDC toabout 50° CA after the TDC (total about 235° CA). The open period 470 ofthe preceding cylinders intake valves is set from about 65° CA beforethe TDC to about the BDC (total about 245° CA). Consequently, a largeoverlap of 115° CA is set from the opening time 472 of the precedingcylinders' intake valves to the closure time 462 of the burnt gasexhaust valves.

If such a large overlap is set, backflow of burnt gas takes place fromthe burnt gas exhaust valves 32 b towards the preceding cylinders'intake valves 31 during the period of this overlap and internal EGR isincreased. Consequently, a fuel costs improvement effect as described inthe first embodiment is obtained and exhaust cleansing is promoted. Itshould be noted that, although increase of the overlap period would tendto invite occurrence of interference of the valve and the upper surfaceof the piston 3, since, in this embodiment, a long stroke configuration(stroke>cylinder bore diameter) is employed, the period for which thepiston 3 is in the vicinity of the top dead center is shortened, therebypreventing such interference.

When, with further increase in engine load, the condition of the regionA502 of FIG. 45 is entered, the open period 460 of the burnt gas exhaustvalves is shifted towards the leading side and the open period 470 ofthe preceding cylinders' intake valves is shifted towards the delayedside by the cam phase varying mechanisms 33 a, 34 a. That is, theoverlap period is shortened and the amount of internal EGR is decreased.Consequently, the ratio of new air is increased, so that the requiredoutput can be obtained.

FIG. 49(b) shows the case of even higher load i.e. the region A503 ofFIG. 45. The upper section thereof shows the open period 500 of theburnt gas exhaust valves and the open period 510 (shown shaded) of thepreceding cylinders' intake valves and the bottom section shows the openperiod 520 of the following cylinders' exhaust valves and the openperiod 530 of the burnt gas introduction valves. These are obtained byshifting the phase of the camshaft 33 by 55° CA in the lagging directionby means of the cam phase varying mechanism 33 a and shifting the phaseof the camshaft 34 by 30° CA in the leading direction by means of thecam phase varying mechanism 34 a. Consequently, the opening time 512 ofthe preceding cylinders' intake valves is about 10° CA before the BDCand the closure time 502 of the burnt gas exhaust valves is about 20° CAafter the BDC and the overlap is set at 30° CA (corresponding to thetypical set value of a conventional engine). The amount of internal EGRis therefore decreased and the ratio of new air is increased, so thedemanded output can be obtained.

Next, a third example of control of intake/exhaust etc using a deviceaccording to this embodiment will be described with reference to FIG.50. The basic construction and basic control configuration of thisexample are the same as in the case of,the second example; the phase ofthe camshafts 33, 34 is varied by the cam phase varying mechanisms 33 a,34 a, without changing over between a low load cam and a high load cam.

FIG. 50(a) shows the case of a comparatively low load region (region A1in FIG. 8) in the operating condition in which the special operatingmode is performed. The upper section thereof shows the open period 540of the burnt gas exhaust valves and the open period 550 (shown shaded)of the preceding cylinders' intake valves and the bottom section showsthe open period 560 of the following cylinders' exhaust valves and theopen period 570 of the burnt gas introduction valves. The open period540 of the burnt gas exhaust valves is set from about 95° CA before theBDC to about 40° CA before the TDC (total about 235° CA). The openperiod 550 of the preceding cylinders' intake valves is set from about10° CA before the TDC to about 55° CA after the BDC (total about 245°CA). Also, the open period 560 of the following cylinders' exhaustvalves is set from about 100° CA before the BDC to about 45° CA beforethe TDC (total about 235° CA). The open period 570 of the burnt gasintroduction valves is set from about 75° CA before the TDC to about 60°CA before the BDC (total about 195° CA).

Considering solely the preceding cylinders 2A, 2D, a minus overlap of30° CA from the closure time 542 of the burnt gas exhaust valves to theopening time 552 of the preceding cylinders' intake valves is thereforeset. However, since the closure time 572 of the burnt gas introductionvalves is set earlier than the closure time 542 of the burnt gas exhaustvalves, burnt gas cannot enter the following cylinders 2B, 2C from thepreceding cylinders 2A, 2D after the closure time 572 of the burnt gasintroduction valves. Consequently, even if the burnt gas exhaust valves32 b are open, the same condition is produced as if they were closed.That is, in fact, the 50° CA from the closure time 572 of the burnt gasintroduction valves to the opening time 552 of the preceding cylinders'intake valves corresponds to a minus overlap. Internal EGR of thepreceding cylinders 2A, 2D is increased by this large minus overlap, soa fuel cost improvement effect as detailed in the first embodiment isobtained and exhaust cleansing is promoted.

When the engine load further increases so that the condition of theregion A502 of FIG. 45 is entered, the open period 540 of the burnt gasexhaust valves is shifted to the delayed side by the cam phase varyingmechanism 34 a. That is, the minus overlap period is shortened and theinternal EGR amount is decreased. Consequently, the ratio of new air isincreased, so that the required output can be obtained.

FIG. 50(b) shows the case where the load becomes even higher so that theregion A503 of FIG. 45 is entered. The upper section shows the openperiod 580 of the burnt gas exhaust valves and the open period 590(shown shaded) of the preceding cylinders' intake valves, while thelower section shows the open period 600 of the following cylinders'exhaust valves and the open period 610 of the burnt gas introductionvalves. These are obtained by shifting the phase of the camshaft 34towards the delayed side with respect to the condition of FIG. 50(a) by60° CA by means of the cam phase varying mechanism 34 a. The openingtime 592 of the preceding cylinders' intake valves therefore becomesabout 10° CA before the BDC and the closure time 582 of the burnt gasexhaust valves becomes about 20° CA after the BDC, so that an overlap of30° CA is set (corresponding to the typical set value of a conventionalengine). The amount of internal EGR is therefore decreased and the ratioof new air is increased, so the required output can be obtained.

Next, control of the intake/exhaust etc using the device of thisembodiment will be described with reference to a fourth example, withreference to FIG. 51. The basic construction and basic controlconfiguration of this embodiment are the same as in the first embodimentbut, depending on conditions, there is partial operation of thefollowing cylinders' intake valves 31 a, which were deactivated in thefirst example in the special operating mode. The mechanism by which thisis produced is partial projection of the peripheral shape of the firstcam 152 a of the cam changeover mechanism 150 a provided for thefollowing cylinders' intake valves 31 a.

FIG. 51(a) is the case of a comparatively low load region (region A501of FIG. 45) of the operating region in which the special operating modeis performed. The upper section thereof shows the open period 620 of theburnt gas exhaust valves and the open period 630 (shown shaded) of thepreceding cylinders' intake valves and the lower section shows the openperiod 644 of the following cylinders' exhaust valves, the open period650 (shown shaded) of the following cylinders' intake valves and theopen period 660 of the burnt gas introduction valves. Directly below theopen period 650 of the following cylinders' intake valves, the openperiod 690 of the following cylinders' intake valves in the ordinaryoperating mode (condition in which the cams are changed over) is shownfor reference. The open period 620 of the burnt gas exhaust valves isset from about 35° CA before the BDC to about 20° CA after the TDC(total about 235° CA). The open period 630 of the preceding cylinders'intake valves is set from about 55° CA before the TDC to about 10° CAafter the BDC (total about 245° CA). Consequently, a large overlap of75° CA from the opening time 632 of the preceding cylinders' intakevalves to the closure time 622 of the burnt gas exhaust valves is set.

If a large overlap is set in this way, the burnt gas flows back from theburnt gas exhaust valves 32 b to the preceding cylinders' intake valves31 during this overlap period, increasing the internal EGR.Consequently, a fuel costs improvement effect as detailed in the firstembodiment is obtained and exhaust cleansing is promoted. It should benoted that, although increase of the overlap period would tend to inviteoccurrence of interference of the valve and the upper surface of thepiston 3, since, in this embodiment, a long stroke configuration(stroke>cylinder bore diameter) is employed, the period for which thepiston 3 is in the vicinity of the top dead center is shortened, therebypreventing such interference.

In addition, new air is arranged to be introduced in the followingcylinders 2B, 2C separately from the burnt gas that is introduced fromthe preceding cylinders 2A, 2D, by providing an open period 650 of thefollowing cylinders' intake valves. The open period 650 of the followingcylinders' intake valves is set from about 65° CA before the TDC toabout 70° CA after the TDC (total about 135° CA). Also, the open period660 of the burnt gas introduction valves is set from about 120° CAbefore the BDC to about 40° CA after the BDC (total about 160° CA).Consequently, new air or burnt gas intake is performed during the periodfrom the opening time 652 of the following cylinders' intake valves tothe closure time 664 of the burnt gas exhaust valves in the followingcylinders 2B, 2C.

Even if the oxygen in the burnt gas that is introduced into thefollowing cylinders 2B, 2C is diminished by increase in the amount ofinternal EGR in the preceding cylinders 2A, 2D, thanks to the provisionof the open period 650 of the following cylinders' intake valves, theoxygen is supplemented by new air separately introduced into thefollowing cylinders 2B, 2C, raising the generated output in thefollowing cylinders 2B, 2C. Also, due to this effect, the limit ofincrease of the amount of internal EGR in the preceding cylinders 2A, 2Dis raised, so the region in which compression self-ignition can beachieved in the preceding cylinders 2A, 2D is further expanded.

Also, since the open time 662 of the burnt gas introduction valves isset on the delayed side from the TDC of the following cylinders 2B, 2Cand the opening time 652 of the following cylinders' intake valves isset earlier than the opening time 662 of the burnt gas introductionvalves and earlier than the TDC of the following cylinders 2B, 2C, theburnt gas that was introduced is prevented from being directly exhaustedthrough the following cylinders' intake valves.

FIG. 51(b) shows the case where the load is even higher and the regionA502 of FIG. 45 is entered. The upper section thereof shows the openperiod 620 of the burnt gas exhaust valves and the open period 670(shown shaded) of the preceding cylinders' intake valves and the bottomsection shows the open period 640 of the following cylinders' exhaustvalves, the open period 680 (shown shaded) of the following cylinders'intake valves and the open period 660 of the burnt gas introductionvalves. These are obtained by shifting the phase of the camshaft 33 inthe lagging direction by means of the cam phase varying mechanism 33 aby 35° CA with respect to the condition of FIG. 51(a). Consequently, theopening time 672 of the preceding cylinders' intake valves becomes about20° CA before the BDC and the opening time 622 of the burnt gas exhaustvalves becomes about 20° CA after the BDC, thereby diminishing theoverlap to 40° CA. The amount of internal EGR is therefore reduced andthe ratio of new air is increased, so the required output at high loadcan be obtained. It should be noted that if the load is furtherincreased, entering the region A503 of FIG. 45, the phase of thecamshaft 33 is further delayed by 10° CA and a changeover to combustionby forced ignition is effected.

It should be noted that instead of dividing the running region A in thespecial operating mode into three regions A501 to A503 it could bedivided into more than these and opening/closure times could be setappropriate to these respective regions. In addition, the changes couldbe effected continuously rather than being set in stepwise fashion bythe divisions. Instead of providing a running region B in the ordinaryoperating mode, the entire region could be treated as running region A.

Although various embodiments of the present invention have beendescribed above, the present invention is not restricted to theseembodiments and could be modified in various ways. Further embodimentsare described below.

(1) Instead of a valve deactivating mechanism as described above, flowpath changeover means could be constituted as in FIG. 52.

Specifically, in this Figure, in the cylinders 2A to 2D of the mainengine body there are respectively provided intake ports 1001 andexhaust ports 1002, intake valves 1003 and exhaust valves 1004 that areprovided at these ports being constantly operated by being opened andclosed under the control of a valve operating mechanism, outside theFigure. Branch intake passages 16A to 16D are connected with the intakeports 1001 of the cylinders 2A to 2D; branch exhaust passages 21A to 21Dare connected with the exhaust ports 1002 of the cylinders 2A to 2D andan inter-cylinder gas passage 1005 is connected between the mergingsection of the branch exhaust passages 21A to 21D with respect to thepreceding cylinders (first and fourth cylinders) 2A and 2D and themerging section of the branch exhaust passages 21B, 21C in respect ofthe following cylinders (second and third cylinders) 2B, 2C and a firstopening/closure valve 1007 is provided in this inter-cylinder gaspassage 1005.

Also, in respect of the preceding cylinders 2A, 2D, the merging sectionof the branch intake passages 16A, 16D is always linked with theupstream section of the intake passages and, in respect of the followingcylinders 2B, 2C, in the linkage section between the merging section ofthe branch intake passages 16B, 16C and the upstream section of theintake passages, a second opening/closure valve 1007 is provided thatopens and closes this linkage section. Furthermore, in respect of thefollowing cylinders 2B, 2C, the merging section of the branch exhaustpassages 21B, 21C is always linked with the downstream section of theexhaust passages and, in respect of the preceding cylinders 2A, 2D, inthe linkage section between the merging section of the branch exhaustpassages 21A, 21D and the downstream section of the exhaust passages, athird opening/closure valve 1008 is provided that opens and closes thislinkage section.

The aforesaid opening/closure valves 1006, 1007 and 1008 are controlledas follows by controller, outside the Figure, depending on whether theoperating condition is in the operating region A on the low load, lowrotational speed side or whether it is in the operating region B on thehigh load or high rotational speed side:

Operating region A: first opening/closure valve 1006 in the opencondition,

second and third opening/closure valves 1007, 1008 in closed condition;and

Operating region B: first opening/closure valve 1006 in the closedcondition,

second and third opening/closure valves 1007, 1008 in open condition.

In this way, a flow path changeover means is constituted by theopening/closure valves 1006, 1007, 1008 and the controller that controlsthese.

A throttle valve 1009 is provided on the upstream side of the mergingsection of the intake passages.

In this embodiment also, in the operating region A, the device is putinto the two-cylinder connected condition, such that, between a pair ofcylinders whose exhaust stroke and intake stroke overlap, the burnt gasexhausted from the preceding cylinders 2A, 2D is directly introducedthrough the inter-cylinder gas passage 1005 to the following cylinders2B, 2C and the gas that is exhausted from these following cylinders 2B,2C is fed to the exhaust passage 20. And in the operating region B, theintake ports 1001 and exhaust ports 1002 of the cylinders 2A to 2D areindependent, so that new air is introduced into the intake ports 1001 ofeach of the cylinders from the intake passages and exhaust gas that isexhausted from the exhaust ports 1002 of each of the cylinders is fed tothe aforementioned exhaust passage 20. The control of fuel injectionfrom the fuel injection valves 9 and the control of ignition etc are thesame as in the basic embodiment.

(2) The device according to the present invention could also be appliedto multi-cylinder engines other than four-cylinder engines. Thus,although, for example in the case of a six-cylinder engine, the exhauststroke of one cylinder cannot fully overlap the intake stroke of anothercylinder, in such cases, it could be arranged for the exhaust stroke ofone cylinder to lead the intake stroke of another cylinder and toconstitute a pair of leading/following cylinders in which the twostrokes of the two cylinders partially overlap.

(3) In addition to a construction as shown in the above embodiments, itcould be arranged for EGR to be performed solely in respect of thepreceding cylinders. If this is done, NOx can be effectively diminished,since production of NOx in the preceding cylinders is suppressed and, inthe following cylinders, the burnt gas introduced from the precedingcylinders suppresses production of NOx in the same way as EGR.

It should be noted that the term “lean air/fuel ratio” has been used andthe word “lean” is mean to be thin but the actual value of air/fuelratio is large.

INDUSTRIAL APPLICABILITY

With the control device according to the present invention, burnt gasexhausted from the exhaustion stroke of the preceding cylinder in a pairof cylinders whose exhaustion stroke and intake stroke overlap isarranged to be directly introduced into the intake stroke of thefollowing cylinder through an inter-cylinder gas passage and gasexhausted from this following cylinder is arranged to be fed to anexhaust passage and combustion is arranged to be performed by forcedignition in the preceding cylinders in a condition with a lean air/fuelratio, while in the following cylinder fuel is supplied to the burnt gasof lean air/fuel ratio introduced from the preceding cylinder andcombustion is arranged to be performed by compression ignition, so fuelcosts can be improved by improvement of thermal efficiency by leancombustion and by decrease in pumping loss in the preceding cylinderwhile, in the following cylinder, the combustion contributes to the workefficiently by performing combustion rapidly by compression ignition;fuel costs can be considerably improved by means of this and lowering ofpumping loss.

In particular, compression ignition can easily be implemented withoutrequiring separate heating means or high pressure compression etc, byutilizing the heat of the burnt gas of high temperature that isintroduced into the following cylinder from the preceding cylinder.Furthermore, since the burnt gas that is introduced into the followingcylinder and the fuel are uniformly distributed, simultaneouscompression ignition can be satisfactorily performed, making it possibleto perform combustion at a high rate and thereby raising thermalefficiency.

1. A control device for a multi-cylinder spark ignition engine havingcylinders arranged to perform a cycle consisting of intake, compression,expansion and exhaustion strokes with prescribed phase differences,characterized in that a gas flow path is constituted in a two-cylinderconnected condition, at least in a low load, low rotational speedregion, such that burnt gas exhausted from a preceding cylinder which isa cylinder on the exhaust stroke side in a pair of cylinders whoseexhaustion stroke and intake stroke overlap is directly introduced intoa following cylinder which is a cylinder on the intake stroke sidethrough an inter-cylinder gas passage and gas exhausted from thisfollowing cylinder is fed to an exhaust passage; and said control devicecomprising combustion controller that controls combustion in eachcylinder such that at least in part of the operating region of theoperating region in which said two-cylinder connected condition isproduced, combustion is performed by forced ignition in said precedingcylinder in a condition at an air/fuel ratio larger by a prescribedamount than the stoichiometric air/fuel ratio while fuel is supplied inan amount corresponding to the following cylinder to the burnt gasgenerated by combustion in this preceding cylinder, and combustion isperformed by compression self-ignition in the following cylinder.
 2. Thecontrol device for a spark ignition engine according to claim 1,characterized in that the air/fuel ratio of the following cylinder insaid two-cylinder connected condition is made to be at or below thestoichiometric air/fuel ratio and a three-way catalyst or oxidationcatalyst is provided in the exhaust passage connected with thisfollowing cylinder.
 3. The control device for a spark ignition engineaccording to claim 1, characterized in that a fuel injection valve isprovided that injects fuel directly into said preceding cylinder and,when in said two-cylinder connected condition, fuel is injected in thecompression stroke from said fuel injection valve and stratified chargecombustion is performed by forced ignition while keeping a lean air/fuelratio in the preceding cylinder.
 4. The control device for a sparkignition engine according to claim 3, characterized in that the air/fuelratio of the preceding cylinder in said two-cylinder connected conditionis twice or more the stoichiometric air/fuel ratio.
 5. The controldevice for a spark ignition engine according to claim 3, characterizedin that the air/fuel ratio of the following cylinder in saidtwo-cylinder connected condition is an air/fuel ratio larger than thestoichiometric air/fuel ratio.
 6. The control device for a sparkignition engine according to any of claims 1 to 5, characterized in thatwhen in said two-cylinder connected condition uniform combustion isperformed by injecting fuel in the following cylinder in the intakestroke.
 7. The control device for a spark ignition engine according toclaim 1, characterized in that it comprises flow path changeover meansfor changing over the flow paths of new air and gas, in a high load,high rotational speed operating region, such that the intake port andexhaust port of each of the cylinders are made to be independent, sothat new air is introduced into the intake port of each cylinder from anintake passage and exhaust gas exhausted from the exhaust port of eachcylinder is fed to said exhaust passage; and combustion controller isarranged to set the air/fuel ratio of each of the cylinders to thestoichiometric air/fuel ratio or less than this and to cause combustionto be performed by forced ignition in each of the cylinders in said highload, high rotational speed operating region.
 8. The control device fora spark ignition engine according to claim 7, characterized in that, insaid preceding cylinder, there are provided an intake port thatcommunicates with said intake passage, a first exhaust port thatcommunicates with said exhaust passage and a second exhaust port thatcommunicates with the inter-cylinder gas passage and, in said followingcylinder there are provided a first intake port that communicates withsaid intake passage, a second intake port that communicates with saidinter-cylinder gas passage and an exhaust port that communicates withsaid exhaust passage; and as said flow path changeover means, there areprovided a valve deactivating mechanism that changes over the operatingcondition and deactivated condition respectively of the first and secondexhaust valves that open and close the first and second exhaust ports ofsaid preceding cylinder and of the first and second intake valves thatopen and close the first and second intake ports of the followingcylinder; and valve stop mechanism controller that, in a low load, lowrotational speed region, puts said first exhaust valve and said firstintake valve in deactivated condition and puts said second exhaust valveand said second intake valve in operating condition and, in a high load,high rotational speed operating condition, puts said first exhaust valveand said first intake valve in operating condition and said secondexhaust valve and said second intake valve in deactivated condition. 9.The control device for a spark ignition four-cycle engine according toclaim 1, characterized in that said combustion controller which exercisethe control mode whereby combustion is performed in said two-cylinderconnected condition as the special operating mode, and said combustioncontroller, in at least part of the operating region of the regioncorresponding to said special operating mode, controls the fuel supplyrate in respect of both the leading and following cylinders such thatthe fuel supply rate in the preceding cylinder is greater, while theair/fuel ratio during combustion in said following cylinder issubstantially the stoichiometric air/fuel ratio, thereby making theair/fuel ratio when combustion is conducted in the preceding cylinder avalue of less than twice the stoichiometric air/fuel ratio andconducting combustion in the preceding cylinder by forced ignition andconducting combustion in the following cylinder by compressionself-ignition.
 10. The control device for a spark ignition four-cycleengine according to claim 9, characterized in that, in said specialoperating mode, in the intermediate speed region of the operating regionin which the following cylinder is made to perform compressionself-ignition, the air/fuel ratio when conducting combustion in thepreceding cylinder is made to be a value of substantially twice thestoichiometric air/fuel ratio, or more than said stoichiometric air/fuelratio.
 11. The control device for a spark ignition four-cycle engineaccording to claim 10, characterized in that, in said special operatingmode, in the operating region on the lower speed side than theintermediate speed region of the operating region in which the followingcylinder is made to perform compression self-ignition, the air/fuelratio when conducting combustion in the preceding cylinder is made to bea value of less than twice the stoichiometric air/fuel ratio.
 12. Thecontrol device for a spark ignition four-cycle engine according to claim10, characterized in that, in said special operating mode, in theoperating region on the higher speed side than the intermediate speedregion of the operating region in which the following cylinder is madeto perform compression self-ignition, the air/fuel ratio when conductingcombustion in the preceding cylinder is made to be a value of less thantwice the stoichiometric air/fuel ratio.
 13. The control device for aspark ignition four-cycle engine according to claim 9, characterized inthat, in said special operating mode, in the intermediate load region ofthe operating region in which the following cylinder is made to performcompression self-ignition, the air/fuel ratio when conducting combustionin the preceding cylinder is made to be a value of substantially twicethe stoichiometric air/fuel ratio, or more than the stoichiometricair/fuel ratio.
 14. The control device for a spark ignition four-cycleengine according to claim 9, characterized in that, in said specialoperating mode, in the intermediate speed/intermediate load region ofthe operating region in which the following cylinder is made to performcompression self-ignition, the air/fuel ratio when conducting combustionin the preceding cylinder is made to be a value of substantially twicethe stoichiometric air/fuel ratio, or more than the stoichiometricair/fuel ratio.
 15. The control device for a spark ignition four-cycleengine according to claim 9, characterized in that, in said specialoperating mode, in the operating region in which the following cylinderis made to perform compression self-ignition, the air/fuel ratio whenconducting combustion in the preceding cylinder is made smaller as theload becomes lower.
 16. The control device for a spark ignitionfour-cycle engine according to claim 9, characterized in that, when theengine temperature is low, in the entire operating region in which thefollowing cylinder is made to perform compression self-ignition in saidspecial operating mode, the air/fuel ratio when conducting combustion inthe preceding cylinder is made to be less than twice the stoichiometricair/fuel ratio.
 17. The control device for a spark ignition engineaccording to claim 1, wherein said combustion controller executescontrol in which combustion is conducted in said two-cylinder connectedcondition in a control mode as a special operating mode; and saidcombustion condition; and said combustion controller including a fuelinjection controller that, in an operating region in which the followingcylinder is made to perform compression self-ignition in said specialoperating mode, relatively retards the injection time of the fuel to thefollowing cylinder in an operating condition in which knocking is likelyto occur, compared with an operating condition where knocking isunlikely to occur.
 18. The control device for a spark ignition engineaccording to claim 17, characterized in that, in an operating region inwhich the following cylinder is made to perform compressionself-ignition in said special operating mode, in an operating conditionin which knocking is likely to occur, the injection time of the fuel tothe following cylinder is set more on the retarded side of thecompression stroke as the likelihood of knocking increases.
 19. Thecontrol device for a spark ignition engine according to claim 17,characterized in that, in an operating region in which the followingcylinder is made to perform compression self-ignition in said specialoperating mode, in an operating condition in which knocking is likely tooccur, injection of fuel into the following cylinder is performed individed fashion and the latter injection time of the fuel in saiddivided injection is set in the latter half of the compression.
 20. Thecontrol device for a spark ignition engine according to claim 19,characterized in that, in a region in which the following cylinder ismade to perform compression self-ignition, the likelihood of occurrenceof knocking or the intensity of knocking is ascertained and the latterinjection time in said divided fuel injection is retarded so as toapproach more closely the compression top dead center as the likelihoodof occurrence of said knocking or the intensity of knocking increases.21. The control device for a spark ignition engine according to claim 3,characterized in that, in a region in which the following cylinder ismade to perform compression self-ignition, in an operating condition inwhich knocking is likely to occur, injection of fuel into the followingcylinder is performed in divided fashion and the latter injection rateof the fuel in said divided injection is set to a larger value than theformer injection rate.
 22. The control device for a spark ignitionengine according to claim 21, characterized in that, in a region inwhich the following cylinder is made to perform compressionself-ignition, the likelihood of occurrence of knocking is ascertainedand the ratio of the latter injection period rate with respect to thetotal injection rate of fuel injected in the following cylinder ischanged so as to be larger as the likelihood of occurrence of suchknocking becomes higher.
 23. The control device for a spark ignitionengine according to claim 17, characterized in that, in a region inwhich the following cylinder is made to perform compressionself-ignition, when the engine is in an operating region on the highload side, a condition in which knocking is likely to occur isidentified.
 24. The control device for a spark ignition engine accordingto claim 17, characterized in that, if fuel of low octane value isemployed, the region in which the following cylinder is made to performcompression self-ignition is identified as a condition in which knockingis likely to occur.
 25. The control device for a spark ignition engineaccording to claim 17, characterized in comprising swirl generatingmeans that generates swirl such that a strong intensity of turbulence ismaintained in the latter half of the compression stroke in a region inwhich the following cylinder is made to perform compressionself-ignition, in an operating condition in which knocking is likely tooccur.
 26. The control device for a spark ignition engine according toclaim 25, characterized in that swirl is generated in the combustionchamber by directing the tip portion of the inter-cylinder gas passagein the cylinder tangential direction of the following cylinder in planview and introducing burnt gas into the following cylinder from saidinter-cylinder gas passage in the intake stroke of the followingcylinder.
 27. The control device for a spark ignition engine accordingto claim 1, characterized in that the combustion controller exercisescontrol whereby combustion is conducted in said two-cylinder connectedcondition as a special operating mode; and said combustion controllereffects a control such that: in at least part of the operating region inwhich said special operating mode is involved, combustion is conductedby compression self-ignition in the following cylinder, and the air/fuelratio of the preceding cylinder is made relatively lower in a high loadregion in the region in which said compression self-ignition isperformed compared with the region on the low load side and a new airintroduction intake valve that introduces new air into the followingcylinder is opened so that new air is introduced into the followingcylinder in addition to the burnt gas that is fed from said precedingcylinder.
 28. The control device for a spark ignition engine accordingto claim 27, characterized in that in a region on the low load side inthe operating region in which the following cylinder is made to performcompression self-ignition in said special operating mode, the new airintroduction intake valve is maintained in closed condition; and, in aregion on the high load side in said compression self-ignition region,the new air introduction intake valve is opened in the vicinity of theintake top dead center of the following cylinder and is closed duringthe course of the intake stroke of the following cylinder.
 29. Thecontrol device for a spark ignition engine according to claim 27,characterized in that, in a region on the high load side in theoperating region in which the following cylinder is made to performcompression self-ignition in said special operating mode, the burnt gasintroduction valve of the following cylinder is opened during the courseof the intake stroke and the new air introduction intake valve is openedprior to the opening time of said burnt gas introduction valve.
 30. Thecontrol device for a spark ignition engine according to claim 27,characterized in that, in a region on the high load side in theoperating region in which the following cylinder is made to performcompression self-ignition, control is exercised such as to increase theratio of the new air intake rate with respect to the total gas rateintroduced into the following cylinder, in response to enrichment of theair/fuel ratio of the preceding cylinder, compared with a region on thelow load side thereof.
 31. The control device for a spark ignitionengine according to claim 27, characterized in that, at least in aregion in which the following cylinder is made to perform compressionself-ignition, the air/fuel ratio of the following cylinder iscontrolled such that the oxygen concentration in the exhaust gas that isexhausted from the following cylinder is a value corresponding to thecombustion condition of the stoichiometric air/fuel ratio.
 32. Thecontrol device for a spark ignition engine according to claim 1,characterized in that control is exercised such as to make the controlmode whereby combustion is conducted in said two-cylinder connectedcondition a special operating mode; and said combustion controllercontrols such that the total injection quantity of fuel injected intothe two cylinders consisting of said preceding cylinder and followingcylinder is increased in response to increase in engine load; andcontrol is exercised such that in said following cylinder, combustion isconducted by compression self-ignition in at least part of the operatingregion in which said special operating mode is involved and, in saidpreceding cylinder, stratified charge lean combustion is conducted withthe injected fuel put in a stratified condition in an intermediate/lowload region of the operating region in which compression self-ignitionof said following cylinder is performed, and control is exercised suchthat, on the high load side of the operating region in which saidstratified charge lean combustion is conducted, uniform lean combustionis conducted in a condition with the injected fuel uniformly dispersed.33. The control device for a spark ignition engine according to claim32, characterized in that, in the operating region on the high load sidein which combustion is conducted in a uniform lean condition in saidpreceding cylinder, the air/fuel ratio of said preceding cylinder ismade to be a value of substantially twice the stoichiometric air/fuelratio, or a value smaller than said stoichiometric air/fuel ratio. 34.The control device for a spark ignition engine according to claim 32,characterized in that, in a low load operating region of theintermediate/low load operating region in which stratified charge leancombustion is conducted in said preceding cylinder, the air/fuel ratioof said preceding cylinder is made to be a value of substantially twicethe stoichiometric air/fuel ratio, or a value smaller than saidstoichiometric air/fuel ratio.
 35. The control device for a sparkignition engine according to claim 32, characterized in that, in a lowload operating region of the intermediate/low load operating region inwhich stratified charge lean combustion is conducted in said precedingcylinder, if compression self-ignition in said following cylinder isdifficult, control is exercised such that the air/fuel ratio of saidpreceding cylinder is made to be substantially twice the stoichiometricair/fuel ratio or a value smaller than said stoichiometric air/fuelratio and the combustion mode in the preceding cylinder is shifted fromthe stratified charge lean condition to said uniform lean condition andthe ignition mode in said following cylinder is shifted from compressionself-ignition to forced ignition.
 36. The control device for a sparkignition engine according to claim 1, characterized in that the flowpaths of intake and exhaust are arranged to be capable of being changedover, these flow paths being capable of being changed over between anordinary operating mode in which each of the cylinders are put in anindependent condition in which combustion is conducted respectivelyindependently and a special operating mode in which combustion isconducted in said two-cylinder connected condition, and said controldevice comprising: first fuel injection means that supplies fuelindependently to each of the cylinders in said ordinary operating mode;second fuel injection means whereby it is made possible to supply fuelin an amount corresponding to that of the following cylinder to saidburnt gas prior to introduction thereof into the following cylinderafter completion of combustion in said preceding cylinder, when in saidspecial operating mode; and wherein said combustion controller, when insaid ordinary operating mode, conducts combustion in at an air/fuelratio in each cylinder, made to be equal to the stoichiometric air/fuelratio by supplying fuel by said first fuel injection means and, when inthe special operating mode, said combustion controller conductscombustion in the preceding cylinder by forced ignition in a conditionat an air/fuel ratio greater by a prescribed amount than thestoichiometric air/fuel ratio, by supplying fuel by said first fuelinjection means, and said combustion controller controls combustion ineach cylinder such as to conduct combustion in the following cylinder bycompression self-ignition by introducing gas in a condition of thestoichiometric air/fuel ratio by supplying fuel to said burnt gas bysaid second fuel injection means.
 37. The control device for a sparkignition engine according to claim 36, characterized in that said firstfuel injection means is arranged such as to inject fuel directly intothe combustion chamber in respect of said preceding cylinder; and thefirst fuel injection means of said preceding cylinder also serves assaid second fuel injection means, when in said special operating mode,by constituting said fuel controller such that supply of fuel for thefollowing cylinder to said burnt gas is performed by said first fuelinjection means of the preceding cylinder during the exhaustion strokeof said cylinder.
 38. The control device for a spark ignition engineaccording to claim 37, characterized in that said first fuel injectionmeans is arranged such that fuel is injected into an intake passage inrespect of said following cylinder.
 39. The control device for a sparkignition engine according to claim 36, characterized in that said secondfuel injection means is provided at some point along said inter-cylindergas passage and fuel is supplied thereby to said burnt gas in an amountcorresponding to that of the following cylinder after exhaustion fromthe preceding cylinder prior to introduction thereof into the followingcylinder.
 40. The control device for a spark ignition engine accordingto claim 36, characterized in that said fuel controller, when in saidspecial operating mode, is capable of changing over the fuel injectionmode between the first injection mode in which combustion is conductedby compression ignition by supplying fuel to said burnt gas in an amountcorresponding to the following cylinder by the first fuel injectionmeans of the following cylinder after introduction of burnt gas into thefollowing cylinder from said preceding cylinder; and a second injectionmode in which combustion is conducted by compression self-ignition bysupplying fuel to said burnt gas in an amount corresponding to thefollowing cylinder by said second fuel injection means prior tointroduction thereof into the following cylinder after completion ofcombustion in said preceding cylinder, and is constituted such as todetermine the degree of capability of self-ignition of the followingcylinder from information relating to the operating condition and to becapable of changing over said injection mode in accordance with theresults of the determination.
 41. The control device for a sparkignition engine according to claim 40, characterized in that saidcombustion controller is constituted such as to put said injection modeinto the second injection mode when in an operating condition whereinthe degree of capability for self-ignition of the following cylinder islow.
 42. The control device for a spark ignition engine according toclaim 41, characterized in that said fuel injection means is constitutedsuch as to determine that the operating condition is the condition inwhich the degree of capability for self-ignition is low if the cylindertemperature is below a specified temperature after warming up operation.43. The control device for a spark ignition engine according to claim41, characterized in that said combustion controller is constituted suchas to determine that the operating condition is one in which the degreeof capability for self-ignition is low when in a very low load region.44. The control device for a spark ignition engine according to claim 1,characterized in that there are provided a preceding cylinder intakevalve for introducing new air into said preceding cylinder and a burntgas introduction valve for introducing burnt gas into said followingcylinder from said inter-cylinder gas passage when in said two-cylinderconnected condition; and in at least a prescribed region on the low loadside of said operating region that is in a two-cylinder connectedcondition, the interval between the intake stroke bottom dead center ofsaid following cylinder and the closure time of said burnt gasintroduction valve is set to be shorter than the interval between theintake stroke bottom dead center of said preceding cylinder and theclosure time of said preceding cylinder intake valve.
 45. The controldevice for a spark ignition engine according to claim 44, characterizedin that there is provided a following cylinder exhaust valve thatexhausts exhaust gas of said following cylinder; and in at least aprescribed region on the low load side of said operating region that isin a two-cylinder connected condition, the opening time of said burntgas introduction valve is set to be the intake stroke top dead center ofsaid following cylinder, while said following cylinder exhaust valve isopen until the top dead center of the exhaust stroke of said followingcylinder.
 46. The control device for a spark ignition engine accordingto claim 44, characterized in that, in a prescribed region on the highload side of said operating region that is in a two-cylinder connectedcondition, the closure time of said burnt gas introduction valve is seton the delayed side from said time when in the prescribed region on thelow load side.
 47. The control device for a spark ignition engineaccording to claim 44, characterized in that, in a prescribed region onthe high load, high rotational speed side of said operating region thatis in a two-cylinder connected condition, the closure time of said burntgas introduction valve is set on the delayed side from said time when inthe prescribed region on the low load, low rotational speed side. 48.The control device for a spark ignition engine according to claim 44,characterized in that a burnt gas exhaust valve is provided thatexhausts burnt gas of said preceding cylinder to said inter-cylinder gaspassage when in said two-cylinder connected condition; and in theoperating region that is in said two-cylinder connected condition, theclosure time of said burnt gas exhaust valve is set on the advancingside of the closure time of said burnt gas introduction valve and whilemaintaining the open period of said burnt gas exhaust valve and the openperiod of said burnt gas introduction valve at fixed prescribed values,the opening time of said burnt gas exhaust valve and the opening time ofsaid burnt gas introduction valve are set so as to vary forwards andbackwards in accordance with engine load while maintaining thedifference of these times fixed.
 49. The control device for a sparkignition engine according to claim 1, characterized in that there areprovided a preceding cylinder intake valve that introduces new air intosaid preceding cylinder and a burnt gas introduction valve thatintroduces burnt gas into said following cylinder from saidinter-cylinder gas passage, when in said two-cylinder connectedcondition; and in at least a prescribed region on the low load side ofthe operating region that is in said two-cylinder connected condition,the open period of said burnt gas introduction valve is set so as to beshorter than the open period of said preceding cylinder intake valve.50. The control device for a spark ignition engine according to claim 1,characterized in that combustion is conducted by compressionself-ignition in said preceding cylinder while increasing the amount ofinternal EGR of said preceding cylinder in a prescribed region on thecomparatively low load side of the operating region in which combustionis conducted by compression self-ignition in the following cylinder andin said two-cylinder connected condition and wherein the internal EGRratio is decreased with increase in load.
 51. The control device for aspark ignition engine according to claim 50, characterized in that, inpart or all of the operating region in which combustion is conducted bycompression self-ignition in both said preceding cylinder and saidfollowing cylinder, the closure time of the burnt gas exhaust valve thatexhausts burnt gas to said inter-cylinder gas passage in the exhauststroke provided in said preceding cylinder is set earlier than the topdead center of the exhaust stroke of said preceding cylinder.
 52. Thecontrol device for a spark ignition engine according to claim 51,characterized in that, in part or all of the operating region in whichcombustion is conducted by compression self-ignition in both saidpreceding cylinder and said following cylinder, said combustioncontroller sets the injection time of fuel into said preceding cylinderlater than the closure time of said burnt gas exhaust valve and in thevicinity of the top dead center of the exhaust stroke.
 53. The controldevice for a spark ignition engine according to claim 51, characterizedin that, in part or all of the operating region in which combustion isconducted by compression self-ignition in both said preceding cylinderand said following cylinder, said combustion controller exercisescontrol such that the air/fuel ratio in said following cylinder issubstantially a lean air/fuel ratio.
 54. The control device for a sparkignition engine according to claim 53, characterized in that thecatalyst provided in said exhaust passage for cleaning exhaust gasconsists solely of a three-way catalyst or solely of a three-waycatalyst and oxidation catalyst.
 55. The control device for a sparkignition engine according to claim 50, characterized in that itcomprises a burnt gas introduction valve provided in said followingcylinder for introducing burnt gas from said inter-cylinder gas passagein the intake stroke when in said two-cylinder connected condition, anda following cylinder intake valve provided in said following cylinderfor introducing new air in the intake stroke when in said two-cylinderconnected condition; and in part all of the operating region in whichcombustion is conducted by compression self-ignition in both saidpreceding cylinder and said following cylinder, the opening time of saidburnt gas introduction valve is set on the delayed side of the top deadcenter of the intake stroke of the following cylinder, and saidfollowing cylinder intake valve is arranged to open earlier than theopening time of said burnt gas introduction valve.
 56. The controldevice for a spark ignition engine according to claim 55, characterizedin that said preceding cylinder is of the long stroke type and in thatit comprises a preceding cylinder intake valve that introduces new airin the intake stroke when in said two-cylinder connected condition; andin part or all of the operating region in which combustion is conductedby compression self-ignition in both said preceding cylinder and saidfollowing cylinder, the closure time of said burnt gas exhaust valve andsaid burnt gas introduction valve is set on the delayed side of the topdead center of the exhaust stroke of said preceding cylinder, and theopening time of said preceding cylinder intake valve is set earlier thanthe top dead center of the intake stroke of the preceding cylinder. 57.The control device for a spark ignition engine according to claim 50,characterized in that it comprises a supercharger that supercharges theintake to said preceding cylinder, and in part or all of the operatingregion in which combustion is conducted by compression self-ignition inat least said preceding cylinder and said following cylinder,supercharging is performed using said supercharger.
 58. The controldevice for a spark ignition engine according to claim 50, characterizedin that, in a prescribed region on the comparatively high load side ofsaid operating region in which combustion is conducted by compressionself-ignition in said following cylinder, said combustion controllerconduct combustion by forced ignition in said preceding cylinder, andset the air/fuel ratio of said preceding cylinder to be substantiallylarger than that when in an operating region in which combustion isconducted by compression self-ignition in both said preceding cylinderand said following cylinder.
 59. A control device for a multi-cylinderspark ignition engine having cylinders arranged to perform a cycleconsisting of intake, compression, expansion and exhaustion strokes withprescribed phase differences, characterized in that a gas flow path isconstituted in a two-cylinder connected condition, at least in a lowload, low rotational speed region, such that burnt gas exhausted from aleading cylinder which is a cylinder on the exhaust stroke side in apair of cylinders whose exhaustion stroke and intake stroke overlap isdirectly introduced into a following cylinder which is a cylinder on theintake stroke side through an inter-cylinder gas passage and gasexhausted from the following cylinder is fed to an exhaust passage; anda three-way catalyst is provided in the exhaust passage connected withthe following cylinder; and said control device comprising combustioncontrol means that controls combustion in each cylinder such that atleast in part of the operating region of the operating region in whichsaid two-cylinder connected condition is produced, combustion isperformed in said leading cylinder in a condition at an air/fuel ratiolarger by a prescribed amount than the stoichiometric air/fuel ratiowhile fuel is supplied in an amount corresponding to the followingcylinder to the burnt gas generated by combustion in the leadingcylinder, and combustion is performed by compression self-ignition atleast in the following cylinder while an amount of fuel injection ineach of said cylinders is controlled in a manner that a total air/fuelratio of both of the leading cylinder and the following cylinder is madeto be larger.
 60. A control device for a four cycled multi-cylinderspark ignition engine having cylinders arranged to perform a cycleconsisting of intake, compression, expansion and exhaustion strokes withprescribed phase differences, and each of said cylinders having anignition plug, characterized in that an inter-cylinder gas passage isprovided between a leading cylinder and a following cylinder in atwo-cylinder connected condition such that burnt gas exhausted from theleading cylinder which is a cylinder on the exhaust stroke side in apair of cylinders whose exhaustion stroke and intake stroke overlap isintroduced into the following cylinder which is a cylinder on the intakestroke side; characterized in that said leading cylinder is providedwith an intake port that communicates with said intake passage, a firstexhaust port that communicates with said exhaust passage and a secondexhaust port that communicates with the inter-cylinder gas passage, andsaid following cylinder is provided with a first intake port thatcommunicates with said intake passage, a second intake port thatcommunicates with said inter-cylinder gas passage and an exhaust portthat communicates with said exhaust passage; characterized in that afirst and a second exhaust valves that open and close the first andsecond exhaust ports of said leading cylinder and a first and a secondintake valves that open and close the first and second intake ports ofthe following cylinder are provided and said first and second exhaustvalves and said first and second intake valves are selectively operatedbetween its activating state and deactivating state, and said controldevice comprising combustion control means that controls a fuelsupplying and an injection in each cylinder in such a manner that: in alow load, low rotational speed region, said first exhaust valve and saidfirst intake valve are set in deactivated condition and said secondexhaust valve and said second intake valve in operating condition sothat the two-cylinder connected condition in which burnt gas exhaustedfrom the leading cylinder which is a cylinder on the exhaust stroke sideis introduced into the following cylinder which is a cylinder on theintake stroke side through an inter-cylinder gas passage is established;characterized in that a three-way catalyst is provided in the exhaustpassage to make the exhaust gas exhausted from the exhaust port of thesaid following cylinder in said two-cylinder connecting condition beingpassed through the three-way catalyst, and when said two-cylinderconnected condition is established, such that combustion is performed insaid leading cylinder at an air/fuel ratio larger by a prescribed amountthan the stoichiometric air/fuel ratio while fuel is supplied in anamount corresponding to the following cylinder to the burnt gasgenerated by combustion in the leading cylinder, and combustion isperformed in said following cylinder at a stoichiometric air/fuel ratioby compression self-ignition.
 61. The control device for a sparkignition four-cycle engine according to claim 60, characterized in thatsaid combustion controller which exercise the control mode wherebycombustion is performed in said two-cylinder connected condition as thespecial operating mode, and said combustion controller, in at least partof the operating region of the region corresponding to said specialoperating mode, controls the fuel supply rate in respect of both theleading and following cylinders such that the fuel supply rate in thepreceding cylinder is greater, while the air/fuel ratio duringcombustion in said following cylinder is substantially thestoichiometric air/fuel ratio, thereby making the air/fuel ratio whencombustion is conducted in the preceding cylinder a value of less thantwice the stoichiometric air/fuel ratio and conducting combustion in thepreceding cylinder by forced ignition and conducting combustion in thefollowing cylinder by compression self-ignition.
 62. The control devicefor a spark ignition engine according to claim 60, wherein saidcombustion controller executes control in which combustion is conductedin said two-cylinder connected condition in a control mode as a specialoperating mode; and said combustion condition; and said combustioncontroller including a fuel injection controller that, in an operatingregion in which the following cylinder is made to perform compressionself-ignition in said special operating mode, relatively retards theinjection time of the fuel to the following cylinder in an operatingcondition in which knocking is likely to occur, compared with anoperating condition where knocking is unlikely to occur.
 63. The controldevice for a spark ignition engine according to claim 60, characterizedin that the combustion controller exercises control whereby combustionis conducted in said two-cylinder connected condition as a specialoperating mode; and said combustion controller effects a control suchthat: in at least part of the operating region in which said specialoperating mode is involved, combustion is conducted by compressionself-ignition in the following cylinder, and the air/fuel ratio of thepreceding cylinder is made relatively lower in a high load region in theregion in which the compression self-ignition is performed compared withthe region on the low load side and a new air introduction intake valvethat introduces new air into the following cylinder is opened so thatnew air is introduced into the following cylinder in addition to theburnt gas that is fed from said preceding cylinder.
 64. The controldevice for a spark ignition engine according to claim 60, characterizedin that control is exercised such as to make the control mode wherebycombustion is conducted in said two-cylinder connected condition aspecial operating mode; and said combustion controller controls suchthat the total injection quantity of fuel injected into the twocylinders consisting of said preceding cylinder and following cylinderis increased in response to increase in engine load; and control isexercised such that in said following cylinder, combustion is conductedby compression self-ignition in at least part of the operating region inwhich said special operating mode is involved and, in said precedingcylinder, stratified charge lean combustion is conducted with theinjected fuel put in a stratified condition in an intermediate/low loadregion of the operating region in which compression self-ignition ofsaid following cylinder is performed, and control is exercised suchthat, on the high load side of the operating region in which thestratified charge lean combustion is conducted, uniform lean combustionis conducted in a condition with the injected fuel uniformly dispersed.65. The control device for a spark ignition engine according to claim60, characterized in that the flow paths of intake and exhaust arearranged to be capable of being changed over, these flow paths beingcapable of being changed over between an ordinary operating mode inwhich each of the cylinders are put in an independent condition in whichcombustion is conducted respectively independently and a specialoperating mode in which combustion is conducted in said two-cylinderconnected condition, and said control device comprising: first fuelinjection means that supplies fuel independently to each of thecylinders in said ordinary operating mode; second fuel injection meanswhereby it is made possible to supply fuel in an amount corresponding tothat of the following cylinder to said burnt gas prior to introductionthereof into the following cylinder after completion of combustion insaid preceding cylinder, when in said special operating mode; andwherein said combustion controller, when in said ordinary operatingmode, conducts combustion in at an air/fuel ratio in each cylinder, madeto be equal to the stoichiometric air/fuel ratio by supplying fuel bysaid first fuel injection means and, when in the special operating mode,said combustion controller conducts combustion in the preceding cylinderby forced ignition in a condition at an air/fuel ratio greater by aprescribed amount than the stoichiometric air/fuel ratio, by supplyingfuel by said first fuel injection means, and said combustion controllercontrols combustion in each cylinder such as to conduct combustion inthe following cylinder by compression self-ignition by introducing gasin a condition of the stoichiometric air/fuel ratio by supplying fuel tosaid burnt gas by said second fuel injection means.
 66. The controldevice for a spark ignition engine according to claim 60, characterizedin that there are provided a preceding cylinder intake valve forintroducing new air into said preceding cylinder and a burnt gasintroduction valve for introducing burnt gas into said followingcylinder from said inter-cylinder gas passage when in said two-cylinderconnected condition; and in at least a prescribed region on the low loadside of said operating region that is in a two-cylinder connectedcondition, the interval between the intake stroke bottom dead center ofsaid following cylinder and the closure time of said burnt gasintroduction valve is set to be shorter than the interval between theintake stroke bottom dead center of said preceding cylinder and theclosure time of said preceding cylinder intake valve.
 67. The controldevice for a spark ignition engine according to claim 60, characterizedin that there are provided a preceding cylinder intake valve thatintroduces new air into said preceding cylinder and a burnt gasintroduction valve that introduces burnt gas into said followingcylinder from said inter-cylinder gas passage, when in said two-cylinderconnected condition; and in at least a prescribed region on the low loadside of the operating region that is in said two-cylinder connectedcondition, the open period of said burnt gas introduction valve is setso as to be shorter than the open period of said preceding cylinderintake valve.
 68. The control device for a spark ignition engineaccording to claim 60, characterized in that combustion is conducted bycompression self-ignition in said preceding cylinder while increasingthe amount of internal EGR of said preceding cylinder in a prescribedregion on the comparatively low load side of the operating region inwhich combustion is conducted by compression self-ignition in thefollowing cylinder and in said two-cylinder connected condition andwherein the internal EGR ratio is decreased with increase in load.
 69. Acontrol device for a multi-cylinder spark ignition engine havingcylinders arranged to perform a cycle consisting of intake, compression,expansion and exhaustion strokes with prescribed phase differences,characterized in that a gas flow path is formed in a two-cylinderconnected condition, at least in a low load, low rotational speedregion, such that burnt gas exhausted from a preceding cylinder which isa cylinder on the exhaust stroke side in a pair of cylinders whoseexhaustion stroke and intake stroke overlap is directly introduced intoa following cylinder which is a cylinder on the intake stroke sidethrough an inter-cylinder gas passage and gas exhausted from thefollowing cylinder is fed to an exhaust passage; and said control devicecomprising a control unit that controls combustion in each cylinder suchthat at least in part of the operating region of the operating region inwhich said two-cylinder connected condition is established, combustionis performed by forced ignition in said preceding cylinder in acondition at an air/fuel ratio larger by a prescribed amount than thestoichiometric air/fuel ratio and fuel is supplied to the followingcylinder in an amount corresponding to the burnt gas generated bycombustion in the preceding cylinder, and combustion is performed bycompression self-ignition in the following cylinder.
 70. The controldevice for a spark ignition four-cycle engine according to claim 69,characterized in that said combustion controller which exercise thecontrol mode whereby combustion is performed in said two-cylinderconnected condition as the special operating mode, and said combustioncontroller, in at least part of the operating region of the regioncorresponding to said special operating mode, controls the fuel supplyrate in respect of both the leading and following cylinders such thatthe fuel supply rate in the preceding cylinder is greater, while theair/fuel ratio during combustion in said following cylinder issubstantially the stoichiometric air/fuel ratio, thereby making theair/fuel ratio when combustion is conducted in the preceding cylinder avalue of less than twice the stoichiometric air/fuel ratio andconducting combustion in the preceding cylinder by forced ignition andconducting combustion in the following cylinder by compressionself-ignition.
 71. The control device for a spark ignition engineaccording to claim 69, wherein said combustion controller executescontrol in which combustion is conducted in said two-cylinder connectedcondition in a control mode as a special operating mode; and saidcombustion condition; and said combustion controller including a fuelinjection controller that, in an operating region in which the followingcylinder is made to perform compression self-ignition in said specialoperating mode, relatively retards the injection time of the fuel to thefollowing cylinder in an operating condition in which knocking is likelyto occur, compared with an operating condition where knocking isunlikely to occur.
 72. The control device for a spark ignition engineaccording to claim 69, characterized in that the combustion controllerexercises control whereby combustion is conducted in said two-cylinderconnected condition as a special operating mode; and said combustioncontroller effects a control such that: in at least part of theoperating region in which said special operating mode is involved,combustion is conducted by compression self-ignition in the followingcylinder, and the air/fuel ratio of the preceding cylinder is maderelatively lower in a high load region in the region in which thecompression self-ignition is performed compared with the region on thelow load side and a new air introduction intake valve that introducesnew air into the following cylinder is opened so that new air isintroduced into the following cylinder in addition to the burnt gas thatis fed from said preceding cylinder.
 73. The control device for a sparkignition engine according to claim 69, characterized in that control isexercised such as to make the control mode whereby combustion isconducted in said two-cylinder connected condition a special operatingmode; and said combustion controller controls such that the totalinjection quantity of fuel injected into the two cylinders consisting ofsaid preceding cylinder and following cylinder is increased in responseto increase in engine load; and control is exercised such that in saidfollowing cylinder, combustion is conducted by compression self-ignitionin at least part of the operating region in which said special operatingmode is involved and, in said preceding cylinder, stratified charge leancombustion is conducted with the injected fuel put in a stratifiedcondition in an intermediate/low load region of the operating region inwhich compression self-ignition of said following cylinder is performed,and control is exercised such that, on the high load side of theoperating region in which the stratified charge lean combustion isconducted, uniform lean combustion is conducted in a condition with theinjected fuel uniformly dispersed.
 74. The control device for a sparkignition engine according to claim 69, characterized in that the flowpaths of intake and exhaust are arranged to be capable of being changedover, these flow paths being capable of being changed over between anordinary operating mode in which each of the cylinders are put in anindependent condition in which combustion is conducted respectivelyindependently and a special operating mode in which combustion isconducted in said two-cylinder connected condition, and said controldevice comprising: first fuel injection means that supplies fuelindependently to each of the cylinders in said ordinary operating mode;second fuel injection means whereby it is made possible to supply fuelin an amount corresponding to that of the following cylinder to saidburnt gas prior to introduction thereof into the following cylinderafter completion of combustion in said preceding cylinder, when in saidspecial operating mode; and wherein said combustion controller, when insaid ordinary operating mode, conducts combustion in at an air/fuelratio in each cylinder, made to be equal to the stoichiometric air/fuelratio by supplying fuel by said first fuel injection means and, when inthe special operating mode, said combustion controller conductscombustion in the preceding cylinder by forced ignition in a conditionat an air/fuel ratio greater by a prescribed amount than thestoichiometric air/fuel ratio, by supplying fuel by said first fuelinjection means, and said combustion controller controls combustion ineach cylinder such as to conduct combustion in the following cylinder bycompression self-ignition by introducing gas in a condition of thestoichiometric air/fuel ratio by supplying fuel to said burnt gas bysaid second fuel injection means.
 75. The control device for a sparkignition engine according to claim 69, characterized in that there areprovided a preceding cylinder intake valve for introducing new air intosaid preceding cylinder and a burnt gas introduction valve forintroducing burnt gas into said following cylinder from saidinter-cylinder gas passage when in said two-cylinder connectedcondition; and in at least a prescribed region on the low load side ofsaid operating region that is in a two-cylinder connected condition, theinterval between the intake stroke bottom dead center of said followingcylinder and the closure time of said burnt gas introduction valve isset to be shorter than the interval between the intake stroke bottomdead center of said preceding cylinder and the closure time of saidpreceding cylinder intake valve.
 76. The control device for a sparkignition engine according to claim 69, characterized in that there areprovided a preceding cylinder intake valve that introduces new air intosaid preceding cylinder and a burnt gas introduction valve thatintroduces burnt gas into said following cylinder from saidinter-cylinder gas passage, when in said two-cylinder connectedcondition; and in at least a prescribed region on the low load side ofthe operating region that is in said two-cylinder connected condition,the open period of said burnt gas introduction valve is set so as to beshorter than the open period of said preceding cylinder intake valve.77. The control device for a spark ignition engine according to claim69, characterized in that combustion is conducted by compressionself-ignition in said preceding cylinder while increasing the amount ofinternal EGR of said preceding cylinder in a prescribed region on thecomparatively low load side of the operating region in which combustionis conducted by compression self-ignition in the following cylinder andin said two-cylinder connected condition and wherein the internal EGRratio is decreased with increase in load.